Refrigerant cycle device

ABSTRACT

A refrigerant cycle device includes a compressor, a using-side heat exchanger which heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and high-pressure refrigerant flowing out of the compressor, an intermediate pressure passage through which intermediate-pressure gas refrigerant obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger is introduced into a intermediate pressure port of the compressor, an exterior heat exchanger which evaporates low-pressure refrigerant obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger and causes the evaporated refrigerant to flow toward a suction port of the compressor, and an auxiliary heater which heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.

CROSS REFERENCE TO RELATED APPLICATION

This application is based on and incorporates herein by reference Japanese Patent Application No. 2011-192559 filed on Sep. 5, 2011 and No. 2012-169404 filed on Jul. 31, 2012.

TECHNICAL FIELD

The present disclosure relates to a vapor-compressing refrigerant cycle device which can be suitably used as a refrigerant cycle device for a vehicle.

BACKGROUND

Conventionally, in Patent Document 1 (JP 7-190574 A), a vehicle air conditioner which performs heating of a vehicle compartment is disclosed. Specifically, air to be blown into the vehicle compartment is heated by using a vapor-compressing refrigerant cycle. In the vehicle air conditioner of Patent Document 1, the blown air, which is a heat-exchange fluid, passes through a using-side heat exchanger and is heated by heat exchange with high-pressure refrigerant discharged from a compressor of the refrigerant cycle.

The vehicle air conditioner further includes an electric heater as an auxiliary heater which is arranged downstream of the using-side heat exchanger in an air flow direction to compensate for a lack of an air heating capacity of the using-side heat exchanger.

Because the vehicle air conditioner includes the electric heater as the auxiliary heater, the electric heater may consume a large amount of electric power in a largest heating operation or the like of the electric heater. Thus, in the largest heating operation or the like, an energy consumption amount of the electric heater may be increased for performing an appropriate heating operation, in which the blown air is heated to a desired temperature.

Furthermore, a heating capacity of the auxiliary heater is required to be so high that the auxiliary heater can compensate for the lack of the air heating capacity of the using-side heat exchanger sufficiently even in a case where the using-side heat exchanger lacks its air heating capacity most. Therefore, the auxiliary heater may become large in size, and the refrigerant cycle device may be thereby increased in size and in manufacturing cost as a whole.

SUMMARY

It is an object of the present disclosure to reduce an energy consumption of a refrigerant cycle device which has a using-side heat exchanger and an auxiliary heater to heat a heat-exchange fluid.

According to an aspect of the present disclosure, a refrigerant cycle device includes a compressor, a using-side heat exchanger, an intermediate pressure passage, an exterior heat exchanger and an auxiliary heater. The compressor has a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion. The using-side heat exchanger heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor. Intermediate-pressure gas refrigerant obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger is introduced into the intermediate pressure port through the intermediate pressure passage. In the exterior heat exchanger, low-pressure refrigerant, obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger, evaporates. The exterior heat exchanger causes the evaporated low-pressure refrigerant to flow to the suction port. The auxiliary heater heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.

Because the refrigerant cycle device includes the auxiliary heater which heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid, an energy consumption of the auxiliary heater for heating the heat-exchange fluid to a target temperature can be reduced as compared with a case where the auxiliary heater heats the heat-exchange fluid that has been heated by the using-side heat exchanger.

Thus, the auxiliary heater increases a temperature of the heat-exchange fluid flowing into the using-side heat exchanger, and a heat radiation amount of refrigerant in the using-side heat exchanger can be thereby decreased. Hence, a cycle balance of the refrigerant cycle can be balanced so that a refrigerant pressure in the using-side heat exchanger increases. Therefore, the above-described reduction of the energy consumption in the auxiliary heater can be achieved.

Accordingly, a temperature of refrigerant discharged from the compressor can be increased, and a temperature difference between refrigerant passing through the using-side heat exchanger and the heat-exchange fluid flowing into the using-side heat exchanger can be thus widen. Furthermore, a compression work amount in a compression process from the intermediate port to the discharge port of the compressor can be increased, and an enthalpy difference between refrigerant flowing at a refrigerant inlet of the using-side heat exchanger and refrigerant flowing at a refrigerant outlet of the using-side heat exchanger can be increased.

Hence, a capacity of the using-side heat exchanger for heating the heat-exchange fluid can be improved. Therefore, the heat-exchange fluid can be heated to the target temperature even though a capacity of the auxiliary heater for heating the heat-exchange fluid is reduced. As a result, the auxiliary heater may have a relatively low heating capacity, and, in this case, an energy consumption of the refrigerant cycle device, in which the using-side heat exchanger and the auxiliary heater are capable of heating the heat-exchange fluid, can be reduced.

According to another aspect of the present disclosure, a refrigerant cycle device includes a compressor, a using-side heat exchanger, a higher-pressure side expansion device, a gas-liquid separation portion, a lower-pressure side expansion device, an exterior heat exchanger and an auxiliary heater. The compressor has a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion. The using-side heat exchanger heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor. The higher-pressure side expansion device is configured to decompress the high-pressure refrigerant flowing out of the using-side heat exchanger into intermediate-pressure refrigerant. The gas-liquid separation portion is configured to separate the intermediate-pressure refrigerant flowing out of the higher-pressure side expansion device into intermediate-pressure gas refrigerant and intermediate-pressure liquid refrigerant. The gas-liquid separation portion causes the separated intermediate-pressure gas refrigerant to flow to the intermediate pressure port. The lower-pressure side expansion device is configured to decompress the separated intermediate-pressure liquid refrigerant flowing out of the gas-liquid separation portion into low-pressure refrigerant. In the exterior heat exchanger, the low-pressure refrigerant flowing out of the lower-pressure side expansion device evaporates. The exterior heat exchanger causes the evaporated low-pressure refrigerant to flow to the suction port. The auxiliary heater heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.

In this case, the refrigerant cycle device is a two stage expansion-type gas injection cycle in which the higher-pressure side expansion device, the gas-liquid separator and the lower-pressure side expansion device are combined. In this case, operational effects similar to those of the aspect of the present disclosure described firstly can be obtained.

According to another aspect of the present disclosure, a refrigerant cycle device includes a compressor, a using-side heat exchanger, a refrigerant branch portion, a first expansion device, an inner heat exchanger, a second expansion device, an exterior heat exchanger and an auxiliary heater. The compressor has a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion. The using-side heat exchanger heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor. At the refrigerant branch portion, a refrigerant passage of the high-pressure refrigerant flowing out of the using-side heat exchanger branches into a first refrigerant passage and a second refrigerant passage. The first expansion device is provided in the first refrigerant passage to decompress the high-pressure refrigerant flowing out of the using-side heat exchanger into intermediate-pressure refrigerant. In the inner heat exchanger, the high-pressure refrigerant flowing from the using-side heat exchanger through the second refrigerant passage exchanges heat with the intermediate-pressure refrigerant decompressed by the first expansion device. The inner heat exchanger causes the heat-exchanged intermediate-pressure refrigerant to flow to the intermediate pressure port. The second expansion device is configured to decompress the heat-exchanged high-pressure refrigerant flowing out of the inner heat exchanger into low-pressure refrigerant. In the exterior heat exchanger, the low-pressure refrigerant flowing out of the second expansion device evaporates. The exterior heat exchanger causes the evaporated low-pressure refrigerant to flow to the suction port. The auxiliary heater heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.

In this case where the refrigerant cycle device is an inner heat-exchange gas injection cycle, operational effects similar to those of the aspect of the present disclosure described firstly can be obtained.

The auxiliary heater may be arranged upstream of the using-side heat exchanger in a flow direction of the heat-exchange fluid to heat the heat-exchange fluid before the heat-exchange fluid being heated by the using-side heat exchanger.

The auxiliary heater and the using-side heat exchanger may be arranged in a direction perpendicular to a flow direction of the heat-exchange fluid, and may be integrated with each other to heat the heat-exchange fluid at the same time.

The auxiliary heater may have a heating capacity lower than a standard heating capacity, which is defined as a necessary heating capacity of the auxiliary heater for heating the heat-exchange fluid to the target temperature in a case where (i) the auxiliary heater is arranged to heat the heat-exchange fluid which has been heated in the using-side heat exchanger, and (ii) the heating capacity of the auxiliary heater and a heating capacity of the using-side heat exchanger are used for heating the heat-exchange fluid.

Because the auxiliary heater has the heating capacity lower than the standard heating capacity, an energy consumption in the auxiliary heater can be reduced reliably as compared with a case where the auxiliary heater heats the heat-exchange fluid after being heated in the using-side heat exchanger. Moreover, the auxiliary heater can be downsized, and the refrigerant cycle device can be thereby reduced in size and in manufacturing cost as a whole.

The refrigerant cycle device may further include a heating capacity adjusting portion configured to adjust the heating capacity of the auxiliary heater so that a pressure of the heat-exchange fluid in the using-side heat exchanger becomes a target pressure.

In this case, because the heating capacity adjusting portion adjusts the heating capacity of the auxiliary heater so that the pressure of the heat-exchange fluid in the using-side heat exchanger becomes the target pressure, the temperature of the heat-exchange fluid can be increased to the target temperature easily by setting the target pressure depending on the target temperature of the heat-exchange fluid.

The auxiliary heater may be an electric heater which generates heat by receiving supply of electric power. In this case, the electric heater, which has a heating capacity lower than the standard heating capacity, includes an electric heater which generates a relatively small heat amount (wattage) when a predetermined voltage is applied to the electric heater.

Alternatively, the auxiliary heater may be an auxiliary heat exchanger which heats the heat-exchange fluid by using heat medium, which cools an external heat source, as a heat source. In this case, the auxiliary heat exchanger, which has a heating capacity lower than the standard heating capacity, includes an auxiliary heat exchanger which has a relatively small area in which the heat-exchange fluid is heated through heat exchange.

When the auxiliary heater is the electric heater as described above, the heating capacity adjusting portion may adjust the heating capacity of the electric heater by adjusting a supply of the electric power to the electric heater.

When the auxiliary heater is the auxiliary heat exchanger as described above, the heating capacity adjusting portion may adjust the heating capacity of the auxiliary heat exchanger by adjusting a flow amount of the heat medium flowing into the auxiliary heat exchanger.

The heating capacity adjusting portion may activate the auxiliary heater when the capacity of the using-side heat exchanger for heating the heat-exchange fluid is incapable of being sufficient by a heating capacity control of the using-side heat exchanger.

In this case, an operation amount of the auxiliary heater can be reduced to requisite minimum.

BRIEF DESCRIPTION OF THE DRAWINGS

The disclosure, together with additional objectives, features and advantages thereof, will be best understood from the following description, the appended claims and the accompanying drawings, in which:

FIG. 1 is a schematic diagram showing a refrigerant circuit of a heat pump cycle for a vehicle air conditioner in a cooling mode and in a dehumidifying-heating mode, according to a first embodiment of the present disclosure;

FIG. 2 is a schematic diagram showing a refrigerant circuit of the heat pump cycle for the vehicle air conditioner in a heating mode, according to the first embodiment;

FIG. 3A is a schematic perspective view showing a gas-liquid separator for the heat pump cycle of the vehicle air conditioner according to the first embodiment;

FIG. 3B is a top view showing the gas-liquid separator for the heat pump cycle of the vehicle air conditioner according to the first embodiment;

FIG. 4 is a flowchart showing a control process of the vehicle air conditioner according to the first embodiment;

FIG. 5 is a flowchart showing a part of the control process of the vehicle air conditioner in the heating mode, according to the first embodiment;

FIG. 6 is a flowchart showing a part of the control process of the vehicle air conditioner in a subcool control of the heating mode, according to the first embodiment;

FIG. 7 is a flowchart showing a part of the control process of the vehicle air conditioner in a quality control (dryness control) of the heating mode, according to the first embodiment;

FIG. 8 is a flowchart showing a part of the control process of the vehicle air conditioner in a PTC-heater control of the heating mode, according to the first embodiment;

FIG. 9 is a flowchart showing a part of the control process of the vehicle air conditioner in the heating mode, according to the first embodiment;

FIG. 10 is a Mollier diagram showing a refrigerant state in the heat pump cycle in the heating mode, according to the first embodiment;

FIG. 11 is a Mollier diagram showing a refrigerant state in a heat pump cycle in a heating mode, according to a comparative example;

FIG. 12 is a schematic diagram showing a refrigerant circuit of a heat pump cycle for a vehicle air conditioner in a heating mode, according to a second embodiment of the present disclosure;

FIG. 13 is a flowchart showing a part of a control process of a vehicle air conditioner in a PTC-heater control of a heating mode, according to a third embodiment of the present disclosure;

FIG. 14 is a schematic diagram showing a refrigerant circuit of a heat pump cycle for a vehicle air conditioner in a heating mode, according to a fourth embodiment of the present disclosure;

FIG. 15 is a Mollier diagram showing a refrigerant state in the heat pump cycle in the heating mode, according to the fourth embodiment; and

FIG. 16 is a schematic diagram showing a refrigerant circuit of a heat pump cycle for a vehicle air conditioner in a heating mode, according to a fifth embodiment of the present disclosure.

DETAILED DESCRIPTION

Embodiments of the present disclosure will be described hereinafter referring to drawings. In the embodiments, a part that corresponds to a matter described in a preceding embodiment may be assigned the same reference numeral, and redundant explanation for the part may be omitted. When only a part of a configuration is described in an embodiment, another preceding embodiment may be applied to the other parts of the configuration. The parts may be combined even if it is not explicitly described that the parts can be combined. The embodiments may be partially combined even if it is not explicitly described that the embodiments can be combined, provided there is no harm in the combination.

First Embodiment

A first embodiment of the present disclosure will be described with reference to FIGS. 1 to 10. In the first embodiment, a refrigerant cycle device of the present disclosure is used for a vehicle air conditioner 1 of an electrical vehicle in which driving force is obtained from an electric motor for vehicle running. In the vehicle air conditioner 1, the refrigerant cycle device functions to heat or cool air to be blown into a vehicle compartment that is a space (air conditioning space) to be air-conditioned. Therefore, the blown air is an example of a fluid (heat-exchange fluid) to be heat-exchanged with refrigerant.

The refrigerant cycle device includes a heat pump cycle 10 (vapor-compressing refrigerant cycle) in which its refrigerant circuit can be switched depending on an air conditioning mode including a cooing mode, a dehumidifying-heating mode (dehumidifying mode) and a heating mode. In the cooling mode, a refrigerant circuit shown in FIG. 1 is selected, and the blown air is cooled to cool the vehicle compartment. Also in the dehumidifying-heating mode, the refrigerant circuit shown in FIG. 1 is selected, and the vehicle compartment is dehumidified and heated. In the heating mode, a refrigerant circuit shown in FIG. 2 is selected, and the blown air is heated to heat the vehicle compartment.

When a hydrofluorocarbon (HFC) refrigerant (e.g., R-134a) is adopted as refrigerant used for the heat pump cycle 10, the heat pump cycle 10 is a vapor-compressing subcritical refrigerant cycle. Thus, a pressure Pd having a highest pressure in the heat pump cycle 10 is lower than a critical pressure of the refrigerant. Alternatively, a hydrofluoro-olefine (HFO) refrigerant (e.g., R1234yf) may be adopted as the refrigerant, for example. The refrigerant contains oil to lubricate a compressor 11 of the heat pump cycle 10, and a part of the oil circulates together with the refrigerant in the heat pump cycle 10.

The compressor 11 of the heat pump cycle 10 is arranged inside a hood of the vehicle, and draws and compresses refrigerant to discharge the compressed refrigerant. The compressor 11 is, for example, an electrical two-stage compressor including a housing used as an outer shell of the compressor 11, higher-stage and lower-stage fixed displacement compression mechanisms accommodated inside the housing, an electric motor accommodated inside the housing to rotationally drive the two compression mechanisms. Refrigerant is compressed at higher pressure in the higher-stage compression mechanism than in the lower-stage compression mechanism.

The housing of the compressor 11 has a suction port 11 a through which low-pressure refrigerant is drawn into the lower-stage compression mechanism from outside the housing, an intermediate pressure port 11 b through which intermediate-pressure refrigerant is drawn into the housing to be mixed with refrigerant flowing from the lower-stage compression mechanism to the higher-stage compression mechanism, and a discharge port 11 c through which high-pressure refrigerant is discharged from the higher-stage compression mechanism to outside the housing.

More specifically, the intermediate pressure port 11 b is connected to a refrigerant discharge side of the lower-stage compression mechanism, in other words, the intermediate pressure port 11 b is connected to a refrigerant, suction side of the higher-stage compression mechanism. Various types of compression mechanisms, such as a scroll-type compression mechanism, a vane-type compression mechanism, and a rolling piston-type compression mechanism, may be adopted as the lower-stage and the higher-stage compression mechanisms.

An operation (rotation rate) of the electric motor of the compressor 11 is controlled by a control signal outputted from an air conditioner controller 40 (A/C ECU), and an alternating-current motor or a direct-current motor may be adopted as the electric motor. By the control of the rotation rate of the electric motor, a refrigerant discharge capacity of the compressor 11 is controlled. Thus, in the present embodiment, the electric motor is used as an example of a discharge capacity changing portion of the compressor 11 which changes the refrigerant discharge capacity of the compressor 11.

The compressor 11 includes the two compression mechanisms accommodated in the single housing of the compressor 11 in the present embodiment, but a configuration of the compressor 11 is not limited to this. Alternatively, the compressor 11 may accommodate a single fixed displacement compression mechanism and an electric motor rotationally driving the single compression mechanism, if intermediate-pressure refrigerant can be drawn into the compressor 11 and can be mixed with refrigerant being in a compression process in the compressor 11.

Moreover, two compressors: higher-stage and lower-stage compressors may be arranged separately in series instead of the above-described configuration of the compressor 11, and the two compressors may be adopted as the single two-stage compressor 11. In this case, a suction port of the lower-stage compressor may be adopted as the suction port 11 a, and a discharge port of the higher-stage compressor may be adopted as the discharge port 11 c. The intermediate pressure port 11 b may be provided in a part connecting a discharge port of the lower-stage compressor and a suction port of the higher-stage compressor.

As shown in FIGS. 1 and 2, the discharge port 11 c of the compressor 11 is connected to a refrigerant inlet side of an interior condenser 12. The interior condenser 12 is arranged inside a casing 31 (air conditioning case) of an interior air conditioning unit 30 of the vehicle air conditioner 1 to function as a radiator in which high-temperature and high-pressure refrigerant discharged from the higher-stage compression mechanism of the compressor 11 radiates heat. The interior condenser 12 is used as an example of a using-side heat exchanger which heats air having passed through an interior evaporator 23 described later.

A refrigerant outlet side of the interior condenser 12 is connected to an inlet of a first expansion valve 13 (higher-stage expansion valve) used as an example of a higher-pressure side expansion device. The higher-pressure side expansion device (13) decompresses high-pressure refrigerant flowing out of the interior condenser 12 so that the high-pressure refrigerant changes into intermediate-pressure refrigerant. The first expansion valve 13 has an electrical variable throttle mechanism. The electrical variable throttle mechanism includes a valve body in which an open degree of the valve body is changeable, and an electrical actuator having a step motor which changes the open degree of the valve body.

When the first expansion valve 13 is set at a decompression state in which the first expansion valve 13 decompresses refrigerant, an open degree of the first expansion valve 13 is regulated within a range from φ0.5 mm to φ3 mm in cross-section diameter. When the first expansion valve 13 is fully open, the open degree is set to be approximately φ10 mm in cross-section diameter. The first expansion valve 13 in the fully open state does not decompress refrigerant. An operation of the first expansion valve 13 is controlled by a control signal outputted from the air conditioning controller 40.

An outlet side of the first expansion valve 13 is connected to an inflow port 14 b of a gas-liquid separator 14. The gas-liquid separator 14 is used as an example of a gas-liquid separation portion which separates intermediate-pressure refrigerant into gas refrigerant and liquid refrigerant. Here, the intermediate-pressure refrigerant has passed through the interior condenser 12 and been compressed in the first expansion valve 13. The gas-liquid separator 14 is a centrifugal separator which separates refrigerant into gas and liquid by utilizing centrifugal force.

A detailed configuration of the gas-liquid separator 14 will be described referring to FIGS. 3A and 3B. The up-down arrow shown in FIG. 3A indicates a vertical direction when the gas-liquid separator 14 is mounted to the vehicle air conditioner 1.

The gas-liquid separator 14 of the present embodiment includes a main body part 14 a, the inflow port 14 b, a gas outflow port 14 c and a liquid outflow port 14 d. The main body part 14 a has a hollow and almost cylindrically bottomed shape with a circular cross-section, and extends in a direction (e.g., the vertical direction) perpendicular to the diameter direction of the circular cross-section. The inflow port 14 b has an inflow opening 14 e through which intermediate-pressure refrigerant is introduced into the main body part 14 a. The gas outflow port 14 c has a gas outflow opening 14 f through which gas refrigerant flows out of the main body part 14 a, and the liquid outflow port 14 d has a liquid outflow opening 14 g through which liquid refrigerant flows out of the main body part 14 a.

A diameter of the main body part 14 a is set at a value from one and half times to three times as large as diameters of refrigerant pipes connected to the ports 14 b to 14 d. Accordingly, the gas-liquid separator 14 is miniaturized.

A volume of the main body part 14 a of the gas-liquid separator 14 is set to be smaller than a surplus refrigerant volume that is obtained by subtracting a necessary refrigerant volume from a sealed total refrigerant volume. Here, the sealed total refrigerant volume is a liquid refrigerant volume converted from a total volume of gas and liquid refrigerant enclosed in the heat pump cycle 10, and the necessary refrigerant volume is a liquid refrigerant volume converted from a necessary volume of refrigerant for optimizing performance of the heat pump cycle 10. In other words, the volume of the gas-liquid separator 14 of the present embodiment is set such that the gas-liquid separator 14 cannot store surplus refrigerant therein substantially, even when a flow rate of refrigerant circulating in the heat pump cycle 10 is changed due to load variation of the heat pump cycle 10.

The inflow port 14 b is connected to a lateral surface of the cylindrical main body part 14 a. As shown in FIG. 3B, the inflow port 14 b extends in a tangential direction of a cross-sectional circle of the main body part 14 a when viewed from above the gas-liquid separator 14. The inflow port 14 b has the inflow opening 14 e at an end of the inflow port 14 b opposite from the main body part 14 a. The inflow port 14 b may not necessarily extend in radial direction (e.g., a horizontal direction), and may extend at some angle with respect to the radial direction.

The gas outflow port 14 c is connected to the main body part 14 a at an upper end surface (top surface) of the main body part 14 a in an axial direction of the main body part 14 a, and the gas outflow port 14 c extends through the top surface of the main body part 14 a coaxially with the main body part 14 a. The gas outflow port 14 c is provided with the gas outflow opening 14 f at an upper end part of the gas outflow port 14 c, and a lower end part of the gas outflow port 14 c is located downward of a connection part between the main body part 14 a and the gas outflow port 14 c.

The liquid outflow port 14 d is connected to the main body part 14 a at a lower end surface (bottom surface) of the main body part 14 a in its axial direction, and the liquid outflow port 14 d extends downward from the bottom surface of the main body part 14 a coaxially with the main body part 14 a. A lower end part of the liquid outflow port 14 d has the liquid outflow opening 14 g.

Refrigerant flowing into the gas-liquid separator 14 from the inflow opening 14 e of the inflow port 14 b flows and swirls along a cylindrical inner surface of the main body part 14 a, and the refrigerant is separated into gas refrigerant and liquid refrigerant by utilizing centrifugal force caused by the swirl flow. Subsequently, the liquid refrigerant obtained by this separation falls down in the main body part 14 a by gravity.

The dropt liquid refrigerant flows out of the liquid outflow opening 14 g of the liquid outflow port 14 d, and the gas refrigerant obtained by the separation flows out of the gas outflow opening 14 f of the gas outflow port 14 c. In FIGS. 3A and 3B, the lower end surface (bottom surface) of the main body part 14 a has a circular shape. Alternatively, the main body part 14 a may be formed into a tapered shape in which a diameter of the main body part 14 a is gradually reduced downward, and a lowest part of the tapered main body part 14 a may be connected to the liquid outflow port 14 d.

As shown in FIGS. 1 and 2, the liquid outflow port 14 c of the gas-liquid separator 14 is coupled to the intermediate pressure port 11 b of the compressor 11 via an intermediate pressure passage 15. A first open-close valve 16 a (intermediate pressure-side open-close valve) is arranged in the intermediate pressure passage 15, and the first open-close valve 16 a is an electromagnetic valve which opens or closes the intermediate pressure passage 15. An operation of the first open-close valve 16 a is controlled by a control signal outputted from the air conditioning controller 40.

The first open-close valve 16 a is used also as a check valve which allows refrigerant only to flow from the gas outflow port 14 c of the gas-liquid separator 14 to the intermediate pressure port 11 b of the compressor 11 when the intermediate pressure passage 15 is open. Accordingly, when the first open-close valve 16 a opens the intermediate pressure passage 15, refrigerant is prevented from flowing back from the compressor 11 to the gas-liquid separator 14.

Moreover, the first open-close valve 16 a functions also to switch the refrigerant circuit of the heat pump cycle 10 by opening or closing the intermediate pressure passage 15. Thus, the first open-close valve 16 a in the present embodiment is used also as an example of a refrigerant circuit switching portion which switches the refrigerant circuit of the heat pump cycle 10.

The liquid outflow port 14 d of the gas-liquid separator 14 is connected to an inlet side of a lower-pressure side fixed throttle 17, and an outlet side of the fixed throttle 17 is connected to a refrigerant inlet side of an exterior heat exchanger 20. The fixed throttle 17 is used as an example of a lower-pressure side expansion device which decompresses liquid refrigerant flowing out of the gas-liquid separator 14 such that a pressure of the liquid refrigerant is reduced to be low-pressure refrigerant. A nozzle having a fixed open degree or an orifice can be adopted as the fixed throttle 17, for example.

In the fixed throttle 17 such as the nozzle or the orifice, a passage cross-section is drastically decreased or drastically increased. Thus, a flow rate of refrigerant flowing through the fixed throttle 17 and a quality (dryness) X of refrigerant upstream of the fixed throttle 17 can be self-adjusted (balanced) depending on a pressure difference between the upstream (inlet) side and a downstream (outlet) side of the fixed throttle 17.

Specifically, when the pressure difference is relatively high, the quality X of refrigerant upstream of the fixed throttle 17 is balanced to be increased in accordance with decrease of a necessary flow amount of refrigerant circulating in the heat pump cycle 10. On the other hand, when the pressure difference is relatively low, the quality X of refrigerant upstream of the fixed throttle 17 is balanced to be decreased in accordance with increase of the necessary flow amount of refrigerant circulating in the heat pump cycle 10.

When the quality X of refrigerant upstream of the fixed throttle 17 is high, and when the exterior heat exchanger 20 is used as an evaporator in which refrigerant is evaporated by absorbing heat, a heat absorption amount (refrigeration capacity) in the exterior heat exchanger 20 may decrease, and a coefficient of performance (COP) of the heat pump cycle 10 may thereby decrease.

Thus, in the present embodiment, the fixed throttle 17 is configured such that the quality X of refrigerant upstream of the fixed throttle 17 is always set to be equal to or lower than 0.1 regardless of change of the necessary flow amount of refrigerant circulating in the heat pump cycle 10 due to the load variation of the heat pump cycle 10 in the heating mode. That is, when a refrigerant circulation rate and the pressure difference between the inlet side and the outlet side of the fixed throttle 17 are changed within an expected range due to the load variation of the heat pump cycle 10, the quality X of refrigerant upstream of the fixed throttle 17 is adjusted to be equal to or lower than 0.1. As a result, the COP of the heat pump cycle 10 can be improved.

The liquid outflow port 14 d of the gas-liquid separator 14 is further connected to a bypass passage 18 through which liquid refrigerant flowing out of the gas-liquid separator 14 bypasses the fixed throttle 17 and is guided toward the exterior heat exchanger 20. A second open-close valve 16 b (low pressure-side open-close valve) is provided in the bypass passage 18. The second open-close valve 16 b is an electromagnetic valve, in which its basic structure is equivalent to a basic structure of the first open-close valve 16 a. An operation of the second open-close valve 16 b is controlled by a control signal outputted from the air conditioner controller 40.

A pressure loss generated when refrigerant flows through the second open-close valve 16 b is extremely lower than a pressure loss generated when refrigerant flows through the fixed throttle 17. Hence, when the second open-close valve 16 b is open, refrigerant from the interior condenser 12 flows into the exterior heat exchanger 20 through the bypass passage 18. On the other hand, when the second open-close valve 16 b is closed, refrigerant from the interior condenser 12 flows into the exterior heat exchanger 20 through the fixed throttle 17.

Thus, the second open-close valve 16 b can cause the refrigerant circuit of the heat pump cycle 10 to be switched. Therefore, the second open-close valve 16 b of the present embodiment is used as an example of the refrigerant circuit switching portion together with the first open-close valve 16 a.

An electrical three-way valve may be used as such refrigerant circuit switching portion (16 b), which switches between a refrigerant circuit connecting an outlet side of the liquid outflow port 14 d of the gas-liquid separator 14 to the inlet side of the fixed throttle 17 and a refrigerant circuit connecting the outlet side of the liquid outflow port 14 d of the gas-liquid separator 14 to an inlet side of the bypass passage 18.

The exterior heat exchanger 20 is arranged in the hood of the vehicle, and refrigerant flowing through the exterior heat exchanger 20 exchanges heat with outside air blown by a blower fan 21. The exterior heat exchanger 20 functions as an evaporator, in which low-pressure refrigerant evaporates and exerts its heat absorption effect in the heating mode, and functions also as a radiator, in which high-pressure refrigerant radiates heat in the cooling mode or the like.

A refrigerant outlet side of the exterior heat exchanger 20 is connected to a refrigerant inlet side of a second expansion valve 22 (cooling expansion valve) which decompresses refrigerant flowing from the exterior heat exchanger 20 to the interior evaporator 23 in the cooling mode or the like. A basic structure of the second expansion valve 22 is similar to that of the first expansion valve 13, and an operation of the second expansion valve 22 is controlled by a control signal outputted from the air conditioning controller 40.

An outlet side of the second expansion valve 22 is connected to a refrigerant inlet side of the interior evaporator 23. The interior evaporator 23 is arranged upstream of the interior condenser 12 in an air flow direction in the casing 31 of the air conditioning unit 30. The interior evaporator 23 is used as an example of an evaporator which cools air by utilizing a heat-absorption effect caused by evaporation of refrigerant flowing through the interior evaporator 23 in the cooling mode, the dehumidifying-heating mode or the like.

A refrigerant outlet side of the interior evaporator 23 is connected to an inlet side of an accumulator 24. The accumulator 24 is a low pressure-side gas-liquid separator which separates refrigerant into gas refrigerant and liquid refrigerant and accumulates surplus refrigerant therein. An outlet of the accumulator 24, through which the gas refrigerant flows out of the accumulator 24, is connected to the suction port 11 a of the compressor 11. The interior evaporator 23 is connected to the suction port 11 a of the compressor 11 via the accumulator 24 such that refrigerant flows from the interior evaporator 23 through the accumulator 24 to the suction port 11 a of the compressor 11.

The refrigerant outlet side of the exterior heat exchanger 20 is further connected to a bypass passage 25, through which refrigerant flowing out of the exterior heat exchanger 20 bypasses the second expansion valve 22 and the interior evaporator 23 to be guided toward the inlet side of the accumulator 24. A third open-close valve 16 c (cooling open-close valve) is provided in the bypass passage 25 to open or close the bypass passage 25.

A basic structure of the third open-close valve 16 c is similar to that of the second open-close valve 16 b, and an operation of the third open-close valve 16 c is controlled by a control signal outputted from the air conditioning controller 40. A pressure loss generated when refrigerant flows through the third open-close valve 16 c is extremely lower than a pressure loss generated when refrigerant flows through the second expansion valve 22.

Hence, when the third open-close valve 16 c is open, refrigerant flowing out of the exterior heat exchanger 20 flows into the accumulator 24 via the bypass passage 25. In this case, the second expansion valve 22 may be fully open.

When the third open-close valve 16 c is closed, refrigerant flowing out of the exterior heat exchanger 20 flows into the interior evaporator 23 via the second expansion valve 22. Therefore, the third open-close valve 16 c can cause the refrigerant circuit of the heat pump cycle 10 to be switched, and the third open-close valve 16 c is used as an example of the refrigerant circuit switching portion together with the first and second open-close valves 16 a, 16 b.

Next, the air conditioning unit 30 will be described with reference to FIGS. 1 and 2. The air conditioning unit 30 is arranged inside an instrumental panel positioned at a front end part of the vehicle compartment. The air conditioning unit 30 includes the casing 31 which constitutes an outer shell of the air conditioning unit 30, and defines therein an air passage through which air is blown toward the vehicle compartment. In the air passage, a blower 32, the interior condenser 12 and the interior evaporator 32 are accommodated, for example.

The casing 31 accommodates an inside/outside air switching device 33, and the inside/outside air switching device 33 is located in an upstream end part of the casing 31. The inside/outside air switching device 33 is used for selectively introducing inside air (REC) (i.e. air inside the vehicle compartment) or/and outside air (FRS) into the casing 31. Specifically, the inside/outside air switching device 33 continuously adjusts an opening area of an inside air port, through which inside air is introduced, and an opening area of an outside air port, through which outside air is introduced, by using an inside/outside air switching door. Accordingly, the inside/outside air switching device 33 continuously changes a ratio between a flow amount of the inside air and a flow amount of the outside air.

The blower 32 is arranged downstream of the inside/outside air switching device 33 in the air flow direction, and the blower 32 blows air, which has been introduced via the inside/outside air switching device 33, toward the vehicle compartment. The blower 32 is an electrical blower which drives a centrifugal multi-blade fan (sirocco fan) by using an electric motor, and a rotation rate (air blowing amount) of the blower 32 is controlled by a control voltage outputted from the air conditioning controller 40.

The interior evaporator 23, a PTC heater 50 (electric heater) and the interior condenser 12 are arranged downstream of the blower 32 in the air flow direction in the order: the interior evaporator 23→the PTC heater 50→the interior condenser 12. In other words, the interior evaporator 23 is arranged upstream of the PTC heater 50 in the air flow direction, and the PTC heater 50 is arranged upstream of the interior condenser 12 in the air flow direction.

The PTC heater 50 is used as an example of an auxiliary heater which heats air in order to compensate for a lack of a capacity of the interior condenser 12 for heating air that is to be blown into the vehicle compartment. More specifically, the PTC heater 50 includes a positive temperature coefficient element (PTC element), and the PTC element receives a supply of electric power from the air conditioning controller 40 to generate heat, thereby heating air that is to flow into the interior condenser 12. A heat generation amount of the PTC heater 50 is increased in accordance with increase of the supplied electric power.

The air conditioning controller 40 of the present embodiment is capable of switching an operation mode of the PTC heater 50. The operation mode of the PTC heater 50 includes a HIGH mode in which the PTC heater 50 receives electric power with a high voltage (e.g., 12V) from the air conditioning controller 40 to output a large heat amount, a LOW mode in which the PTC heater 50 receives electric power with a low voltage (e.g., 6V) from the air conditioning controller 40 to output a small heat amount, and an OFF mode in which the PTC heater 50 is not energized.

Here, an air heating capacity of the PTC heater 50 of the present embodiment will be described. The inventor has studied a case where a PTC heater is arranged so as to heat air which has been heated in the interior condenser 12. In other words, the PTC heater is arranged downstream of the interior condenser 12 instead of the PTC heater 50 arranged upstream of the interior condenser 12. Then, air is heated to a target temperature by using heating capacities of both the interior condenser 12 and the PTC heater.

In this case, when the air heating capacity of the interior condenser 12 is smallest, a necessary heating capacity (largest heating capacity) of the PTC heater for heating the air to the target temperature is approximately 2kW. In other words, the PTC heater is required to output heat of 2kW when the PTC heater is energized with a rated voltage of 12V, for example. The largest hating capacity of the PTC heater is referred to as a standard heating capacity hereinafter.

In the present embodiment, the, PTC heater 50 has an air heating capacity lower than the standard heating capacity. For example, the PTC heater 50 generates heat of approximately 800W, which is equal to or smaller than a half of the standard heating capacity, when the PTC heater 50 is energized with a rated voltage of 12V.

In the casing 31, a bypass air passage 35 is provided, through which air having passed through the interior evaporator 23 bypasses the interior condenser 12 and the PTC heater 50, and an air mix door 34 is arranged downstream of the interior evaporator 23 and upstream of the interior condenser 12 in the air flow direction.

The air mix door 34 of the present embodiment adjusts a ratio between a flow amount of air, which passes through the PTC heater 50 and the interior condenser 12, and a flow amount of air, which passes through the bypass air passage 35. Hence, the air mix door 34 is used as an example of a flow amount adjusting portion which adjusts a flow amount (air amount) flowing into the interior condenser 12, and is used also as an example of a heat exchange capacity adjusting portion which adjusts a heat exchange capacity of the interior condenser 12.

A mixing space 36 is provided downstream of the interior condenser 12 and of the bypass air passage 35 in the air flow direction in the casing 31. Heated air, which has exchanged heat with refrigerant in the interior condenser 12, and non-heated air, which has passed through the bypass air passage 35, are mixed with each other in the mixing space 36. The mixing space 36 is used as an air mix chamber in which heated air (warm air) and non-heated air (cool air) are mixed with each other.

Thus, the air mix door 34 adjusts the ratio between the flow amount of air passing through the interior condenser 12 and the flow amount of air passing through the bypass air passage 35, so that a temperature of air in the mixing space 36 is adjusted. The air mix door 34 is driven by a servomotor, and an operation of the servomotor is controlled by a control signal outputted from the air conditioning controller 40.

A downstream part of the casing 31 in the air flow direction has air outlet openings through which air conditioned in the mixing space 36 is blown out toward the vehicle compartment, and the vehicle compartment is the space (air conditioning space) to be air-conditioned. The air outlet openings include a defroster opening 37 a through which conditioned air is blown toward an inner surface of a windshield of the vehicle, a face opening 37 b through which conditioned air is blown toward an upper part of a passenger in the vehicle compartment, and a foot openings 37 c through which conditioned air is blown toward a foot area of the passenger.

A defroster door 38 a, a face door 38 b and a foot door 38 c are arranged upstream of the defroster opening 37 a, the face opening 37 b and the foot opening 37 c in the air flow direction, respectively. The defroster door 38 a, the face door 38 b and the foot door 38 c adjust opening areas of the defroster opening 37 a, the face opening 37 b and the foot opening 37 c, respectively.

The defroster door 38 a, the face door 38 b and the foot door 38 c open or close their openings 37 a, 37 b, 37 c, respectively, thereby being used as examples of an air outlet mode switching portion which switches an air outlet mode. These three doors 38 a, 38 b, 38 c are driven via a link mechanism or the like by a servomotor in which an operation of the servomotor is controlled by a control signal outputted from the air conditioning controller 40.

Downstream sides of the defroster opening 37 a, the face opening 37 b and the foot opening 37 c in the air flow direction are connected respectively to a defroster air outlet, a face air outlet and a foot air outlet through air passages of ducts. The defroster air outlet, the face air outlet and the foot air outlet are provided in the vehicle compartment.

The air outlet mode includes a face mode in which the face opening 37 b is fully open such that conditioned air is blown from the face air outlet toward the upper part of the passenger in the vehicle compartment, a bi-level mode in which both the face opening 37 b and the foot opening 37 c are open such that conditioned air is blown toward the upper part and the foot area of the passenger, and a foot mode in which the foot opening 37 c is fully open and the defroster opening 37 a is slightly open such that conditioned air is blown mainly from the foot air outlet.

Next, an electrical control portion of the present embodiment will be described. The air conditioning controller 40 includes a known microcomputer and its peripheral circuit, and the microcomputer includes a central processing unit (CPU), a read-only memory (ROM) and a random access memory (RAM). The air conditioning controller 40 performs various calculation and processes based on an air conditioning program stored in the ROM, and controls operations of various air conditioning components (e.g., the compressor 11, the open-close valves 16 a, 16 b, 16 c, the blower 32 and the PTC heater 50) connected to an output side of the air conditioning controller 40.

An input side of the air conditioning controller 40 is connected to a sensor group 41 having various air conditioning sensors. The sensor group 41 includes an inside air sensor which detects a temperature inside the vehicle compartment, an outside air sensor which detects a temperature of outside air, a solar radiation sensor which detects a solar radiation amount entering into the vehicle compartment, an evaporator temperature sensor which detects a temperature (temperature of the interior evaporator 23) of air flowing out of the interior evaporator 23, a discharge pressure sensor which detects a pressure of high-pressure refrigerant discharged from the compressor 11, a condenser temperature sensor which detects a temperature of refrigerant flowing out of the interior condenser 12, and an inlet pressure sensor which detects a pressure of refrigerant drawn into the compressor 11.

The input side of the air conditioning controller 40 is further connected to a control panel (not shown) disposed in the instrumental panel at the front end part of the vehicle compartment, and control signals from various air conditioning switches provided on the control panel are inputted into the air conditioning controller 40. The various air conditioning switches of the control panel includes an activation switch of the vehicle air conditioner 1, a temperature setting switch used for setting a temperature inside the vehicle compartment, and a mode selecting switch used for selecting the air conditioning mode from among the cooling mode, the dehumidifying-heating mode and the heating mode.

The air conditioning controller 40 is further connected to a non-shown battery which outputs a rated voltage of 12V, and the battery supplies electric power to the air conditioning controller 40. The air conditioning controller 40 is capable of transforming the supplied electric power, and thereby supplies the transformed electric power to the various air conditioning components such as the PTC heater 50.

The air conditioning controller 40 integrally includes control portions (hardware and software) which respectively control operations of the various air conditioning components connected the output side of the air conditioning controller 40.

For example, in the present embodiment, the control portions include a discharge capacity control portion which controls an operation of the electric motor of the compressor 11, a refrigerant circuit control portion which controls operations of the open-close valves 16 a, 16 b, 16 c used as examples of the refrigerant circuit switching portion, and a heating capacity control portion 40 a used as an example of a heating capacity adjusting portion which adjusts the air heating capacity of the auxiliary heater (e.g., PTC heater 50) by adjusting an electric energy supplied to the auxiliary heater. The discharge capacity control portion, the refrigerant circuit control portion and the heating capacity control portion 40 a may be provided separately from the air conditioning controller 40.

Next, an operation of the vehicle air conditioner 1 of the present embodiment will be described in reference to FIGS. 4 to 9. A control process of the operation of the vehicle air conditioner 1 shown in FIG. 4 starts when the activation switch of the vehicle air conditioner 1 is turned ON. Each control step of flowcharts shown in drawings constitutes each of a variety of function execution portions that the air conditioning controller 40 includes.

At step S1, the air conditioning controller 40 performs initializations (initializing process) of a flag, a timer, default positions of the above-described various electrical actuators and the like, and then performs a control operation of step S2. In the initializing process of step S1, some of flags and calculation values stored at termination of the last operation of the vehicle air conditioner 1 are maintained.

At step S2, the air conditioning controller 40 reads in control signals from the control panel, such as a preset temperature Tset of the vehicle compartment set by the temperature setting switch and an air conditioning mode selected by the mode selecting switch. Subsequently, a control operation of step S3 is performed. At step S3, the air conditioning controller 40 reads in signals of vehicle environmental conditions used for performing an air conditioning control. In other words, the air conditioning controller 40 reads in detection signals from the sensor group 41 for performing the air conditioning control. Then, the air conditioning controller 40 performs a control operation of step S4.

At step S54, the air conditioning controller 40 calculates a target outlet temperature TAO (target temperature) of air to be blown into the vehicle compartment from the air outlets, and then performs a control operation of step S5. Specifically, at step S4, the target outlet temperature TAO of the present embodiment is calculated by using the preset temperature Tset, an inside air temperature Tr of the vehicle compartment detected by the inside air sensor, an outside air temperature Tam detected by the outside air sensor, and a solar radiation amount Ts detected by the solar radiation sensor.

At step S5, the air conditioning controller 40 determines an air blowing capacity (air blowing amount) of the blower 32, and then performs a control operation of step S6. Specifically, at step S5, the air blowing amount (e.g., a blower motor voltage applied to the electric motor of the blower 32) of the blower 32 is determined by using a control map stored in the air conditioning controller 40 based on the target outlet temperature TAO determined at step S4.

For example, in the present embodiment, when the target outlet temperature TAO is determined within an extremely low temperature range or an extremely high temperature range at step S4, the blower motor voltage is set to be high voltage around a highest value so that the air blowing amount of the blower 32 is controlled to be around a largest air blowing amount. When the target outlet temperature TAO is increased from the extremely low temperature range toward a middle temperature range, the blower motor voltage is reduced so that the air blowing amount of the blower 32 is reduced in accordance with the increase of the target outlet temperature TAO.

When the target outlet temperature TAO is decreased from the extremely high temperature range toward the middle temperature range, the blower motor voltage is reduced so that the air blowing amount of the blower 32 is reduced in accordance with the decrease of the target outlet temperature TAO. When the TAO is determined to be within the middle temperature range, the blower motor voltage is set to be a lowest value so that the air blowing amount of the blower 32 becomes a smallest amount.

At step S6, the air conditioning controller 40 determines the air conditioning mode based on a control signal inputted from the mode selecting switch of the control panel. When the cooling mode is selected as the air conditioning mode by the mode selecting switch, a control process of step S7 is performed. When the dehumidifying-heating mode is selected as the air conditioning mode, a control process of step S8 is performed. When the heating mode is selected as the air conditioning mode, a control process of step S9 is performed.

At steps S7 to S9, the control processes corresponding to each air conditioning mode are performed, and then a control operation of step S10 is performed. Details of the control processes of steps S7 to S9 will be described later.

At step S10, the air conditioning controller 40 determines a switching condition (air inlet mode) of the inside/outside air switching device 33, and then performs a control operation of step S11. At step S10, the air inlet mode is determined based on the target outlet temperature TAO by using a control map stored in the air conditioning controller 40. In the present embodiment, an outside air mode, in which outside air is mainly introduced into the air conditioning unit 30, is generally determined as the air inlet mode. However, when the target outlet temperature TAO is determined to be within the extremely low temperature range or within the extremely high temperature range, in other words, when high cooling performance or high heating performance is required, an inside air mode is selected as the air inlet mode, in which inside air is mainly introduced into the air conditioning unit 30.

At step S11, the air conditioning controller 40 determines the air outlet mode, and then performs a control operation of step S12. At step S11, the air outlet mode is determined based on the target outlet temperature TAO by using a control map stored in the air conditioning controller 40. In the present embodiment, the air outlet mode is switched in an order: the foot mode→the bi-level mode→the face mode, in accordance with change of the target outlet temperature TAO from a high temperature range to a low temperature range.

At step S12, the air conditioning controller 40 outputs control signals and control voltages to the various air conditioning components, which are connected to the output side of the air conditioning controller 40 to be controlled, such that control states determined at steps S6 to S11 are obtained. At step S13, the air conditioning controller 40 waits for a control period τ. The air conditioning controller 40 determines the elapse of the control period τ, and then performs the control operation of step S2.

In the control routine shown in FIG. 4, the air conditioning controller 40 repeats the above-described control operations: reading detection signals and control signals→determination of the control states of the various controlled components→output of control signals and control voltages to the various controlled components. The control routine is performed until the operation of the vehicle air conditioner 1 is required to be stopped by turning the activation switch OFF, for example. Next, details of the air conditioning modes performed at steps S7 to S9 will be described.

(a) Cooling Mode

The cooling mode performed at step S7 will be described. In the cooling mode, the air conditioning controller 40 fully opens the first expansion valve 13, and makes the second expansion valve 22 to be in a decompression state in which its open degree is reduced and its decompression effect is exerted. Additionally, the air conditioning controller 40 closes the first and third open-close valves 16 a, 16 c, and opens the second open-close valve 16 b.

Thus, when the air conditioning controller 40 outputs control signals and control voltages to the various controlled components at step S12 shown in FIG. 4, the refrigerant circuit of the heat pump cycle 10 as shown by solid arrows in FIG. 1 is provided. In this cycle configuration of the cooling mode, the air conditioning controller 40 determines operation states of the various air conditioning components connected to the output side of the air conditioning controller 40 based on the target outlet temperature TAO determined at step S4 and detection signals inputted from the sensor group 41.

For example, a rotation rate Nc of the compressor 11 (i.e., control signal outputted to the electric motor of the compressor 11) is determined as follows. First, a target evaporator temperature TEO of the interior evaporator 23 is determined based on the target outlet temperature TAO by using a control map stored in the air conditioning controller 40. The target evaporator temperature TEO is determined so as to be equal to or higher than a predetermined temperature (e.g., 1° C.) which is higher than a frost formation temperature (i.e., 0° C.), in order to prevent the interior evaporator 23 from frosting.

And then, the rotation rate Nc is determined based on a deviation between the target evaporator temperature TEO and a temperature of air flowing out of the interior evaporator 23 detected by the evaporator temperature sensor, so that the temperature of air flowing out of the interior evaporator 23 approaches the target evaporator temperature TEO by a feedback control.

A control signal outputted to the second expansion valve 22 is determined such that a supercooling degree of refrigerant flowing into the second expansion valve 22 approaches a predetermined target supercooling degree. The target supercooling degree is determined so that the COP approaches approximately a largest value. A control signal outputted to the servomotor of the air mix door 34 is determined so that the air mix door 34 closes an air passage of the interior condenser 12, and so that a total amount of air flowing out of the interior evaporator 23 flows into the bypass air passage 35.

The above-described control routine such as, reading detection signals and control signals→calculation of the target outlet temperature TAO→determination of the operation conditions of the various air conditioning components→output of control voltages and control signals, is repeated until the air conditioning mode is switched to the dehumidifying-heating mode or to the heating mode at step S6 shown in FIG. 4, or until the vehicle air conditioner 1 is required to be stopped by, for example, a control signal from the control panel.

In the cooling mode of the heat pump cycle 10, high-pressure refrigerant discharged from the discharge port 11 c of the compressor 11 flows into the interior condenser 12. Because the air mix door 34 closes the air passage of the interior condenser 12, the high-pressure refrigerant flows through the interior condenser 12 with radiating heat little.

The high-pressure refrigerant flowing out of the interior condenser 12 flows through in an order: the first expansion valve 13→the gas-liquid separator 14→the second open-close valve 16 b, and then flows into the exterior heat exchanger 20. Because the first expansion valve 13 is fully open, the high-pressure refrigerant flowing out of the interior condenser 12 flows through the first expansion valve 13 with little decompression. Subsequently, the refrigerant flowing out of the first expansion valve 13 flows into the gas-liquid separator 14 from the inflow port 14 b of the gas-liquid separator 14.

Here, the refrigerant flowing into the gas-liquid separator 14 is in a gas state because the refrigerant has exchanged heat little with air in the interior condenser 12. Thus, the gas refrigerant flows out of the liquid outflow port 14 d without gas-liquid separation in the gas-liquid separator 14. Moreover, the first open-close valve 16 a is closed, so that the gas refrigerant does not flow out of the gas outflow port 14 c.

The high-pressure gas refrigerant flowing out of the liquid outflow port 14 d flows into the exterior heat exchanger 20 via the bypass passage 18 without flowing into the fixed throttle 17 because the second open-close valve 16 b is open. The high-pressure refrigerant flowing into the exterior heat exchanger 20 radiates heat through heat exchange with outside air blown by the blower fan 21, and thereby condenses.

The refrigerant flowing out of the exterior heat exchanger 20 flows into the second expansion valve 22 which is in the decompression state, because the third open-close valve 16 c is closed. Then, the refrigerant flowing into the second expansion valve 22 changes into low-pressure refrigerant through isenthalpic expansion and decompression. The low-pressure refrigerant having decompressed in the second expansion valve 22 flows into the interior evaporator 23 to absorb heat from air blown by the blower 32 and to evaporate. Accordingly, the air to be blown into the vehicle compartment is cooled.

The refrigerant flowing out of the interior evaporator 23 is separated into gas refrigerant and liquid refrigerant in the accumulator 24. The gas refrigerant is drawn into the compressor 11 from the suction port 11 a, and is compressed again by the lower-stage compression mechanism and then by the higher-stage compression mechanism. As described above, in the cooling mode, because the air mix door 34 closes the air passage of the interior condenser 12, air cooled by the interior evaporator 23 can be blown into the vehicle compartment with the air kept in a cool state. In other words, the air cooled by the interior evaporator 23 can be blown into the vehicle compartment without passing through the interior condenser 12. Accordingly, cooling of the vehicle compartment can be performed.

(b) Dehumidifying-Heating Mode

Details of the dehumidifying-heating mode performed at step S8 will be described. In the dehumidifying-heating mode, the air conditioning controller 40 makes the first expansion valve 13 to be in a fully open state or in the decompression state, and makes the second expansion valve 22 to be in a fully open state or in the decompression state. Moreover, the air conditioning controller 40 closes the first and third open-close valves 16 a, 16 c, and opens the second open-close valve 16 b. Therefore, a refrigerant circuit of the heat pump cycle 10 shown by the solid arrows in FIG. 1, which is similar to the refrigerant circuit of the heat pump cycle 10 in the cooling mode, is provided.

Additionally, the air conditioning controller 40 determines a control signal that is to be outputted to the servomotor of the air mix door 34 such that an open degree of the air mix door 34 is set to be smallest to close the bypass air passage 35. Accordingly, an entire flow amount of air having passed through the interior evaporator 23 flows through the interior condenser 12. However, the open degree of the air mix door 34 may be adjusted based on the target outlet temperature TAO, even in the dehumidifying-heating mode.

The rotation rate Nc of the compressor 11 is determined such that the higher-pressure side refrigerant pressure Pd between the discharge port 11 c of the compressor 11 and the inlet side of the first expansion valve 13 in the heat pump cycle 10 approaches the target pressure TPd by a feedback control or the like. The target pressure TPd is determined based on the target outlet temperature TAO by using a control map stored in the air conditioning controller 40, such that the temperature of air to be blown into the vehicle compartment becomes the target outlet temperature TAO.

In the dehumidifying-heating mode of the present embodiment, open degrees of the first and second expansion valves 13, 22 are changed depending on a temperature difference between the preset temperature Tset and the outside air temperature Tam. Specifically, the dehumidifying-heating mode includes first to forth dehumidifying-heating modes, and one of the four dehumidifying-heating modes is performed in the dehumidifying-heating mode depending on the target outlet temperature TAO.

(b)(i) First Dehumidifying-Heating Mode

A first dehumidifying-heating mode is one example of the dehumidifying-heating mode. In the first dehumidifying-heating mode, the first expansion valve 13 is fully open and the second expansion valve 22 is in a decompression state, so that a cycle configuration (refrigerant circuit) of the first dehumidifying-heating mode is similar to that of the cooling mode. The air mix door 34 is adjusted to fully open the air passage of the interior condenser 12.

High-pressure refrigerant discharged from the discharge port 11 c of the compressor 11 flows into the interior condenser 12. Then, the high-pressure refrigerant radiates heat and condenses through heat exchange with air that has been cooled and dehumidified by the interior evaporator 23. Accordingly, air to be blown into the vehicle compartment is heated by the interior condenser 12. The refrigerant flowing out of the interior condenser 12 flows through in the order: the first expansion valve 13→the gas-liquid separator 14→the second open-close valve 16 b, and then the refrigerant flows into the exterior heat exchanger 20. The high-pressure refrigerant flowing into the heat exchanger 20 radiates heat and condenses through heat exchange with outside air blown by the blower fan 21. A subsequent refrigerant flow and corresponding state change of the refrigerant are similar to those of the cooling mode. That is, the other operation states of the first dehumidifying-heating mode are similar to those of the cooling mode.

As described above, in the first dehumidifying-heating mode, air having been cooled and dehumidified in the interior evaporator 23 can be heated in the interior condenser 12 and can be blown into the vehicle compartment. Accordingly, dehumidifying and heating of the vehicle compartment can be performed.

(b)(ii) Second Dehumidifying-Heating Mode

When the target outlet temperature TAO becomes higher than a first reference temperature during the operation of the first dehumidifying-heating mode, a second dehumidifying-heating mode is performed. The second dehumidifying-heating mode is another example of the dehumidifying-heating mode. In the second dehumidifying-heating mode, the first expansion valve 13 is in a decompression state, and the second expansion valve 22 is in a decompression state in which the open degree of the second expansion valve 22 is larger than that in the first dehumidifying-heating mode.

Similarly to the first dehumidifying-heating mode, high-pressure refrigerant discharged from the discharge port 11 c of the compressor 11 flows into the interior condenser 12 and radiates heat by heat exchange with air having been cooled and dehumidified in the interior evaporator 23. Accordingly, air to be blown into the vehicle compartment is heated in the interior condenser 12.

The high-pressure refrigerant flowing out of the interior condenser 12 changes into intermediate-pressure refrigerant through isenthalpic decompression in the first expansion valve 13 which is in a decompression state. The intermediate-pressure refrigerant flowing out of the first expansion valve 13 flows through in an order: the gas-liquid separator 14→the second open-close valve 16 b, and then the refrigerant flows into the exterior heat exchanger 20. The intermediate-pressure refrigerant flowing into the exterior heat exchanger 20 radiates heat by heat exchange with outside air blown by the blower fan 21. A subsequent refrigerant flow and corresponding state change of the refrigerant are similar to those of the cooling mode.

As described above, in the second dehumidifying-heating mode, air having been cooled and dehumidified in the interior evaporator 23 can be heated in the interior condenser 12 and can be blown into the vehicle compartment, similarly to the first dehumidifying-heating mode. Accordingly, dehumidifying and heating of the vehicle compartment can be performed.

Because the first expansion valve 13 is in the decompression state in the second dehumidifying-heating mode, a temperature of refrigerant passing through the exterior heat exchanger 20 can be reduced relative to the case of the first dehumidifying-heating mode. Thus, a temperature difference between the refrigerant and outside air in the exterior heat exchanger 20 can be reduced, and a heat radiation amount of refrigerant in the exterior heat exchanger 20 can be thereby reduced.

As a result, in the second dehumidifying-heating mode, a heat radiation amount of refrigerant in the interior condenser 12 can be increased, and the air heating capacity of the interior condenser 12 can be thereby improved relative to the case of the first dehumidifying-heating mode.

(b)(iii) Third Dehumidifying-Heating Mode

When the target outlet temperature TAO becomes higher than a second reference temperature during the operation of the second dehumidifying-heating mode, a third dehumidifying-heating mode is performed. The third dehumidifying-heating mode is another example of the dehumidifying-heating mode. In the third dehumidifying-heating mode, the open degree of the first expansion valve 13 is adjusted to be smaller than that in the second dehumidifying-heating mode, and the open degree of the second expansion valve 22 is adjusted to be larger than that in the second dehumidifying-heating mode.

Similarly to the first and second dehumidifying-heating modes, high-pressure refrigerant discharged from the discharge port 11 c of the compressor 11 flows into the interior condenser 12 and radiates heat by heat exchange with air having been cooled and dehumidified in the interior evaporator 23 in the third dehumidifying-heating mode. Accordingly, air to be blown into the vehicle compartment is heated in the interior condenser 12.

The high-pressure refrigerant flowing out of the interior condenser 12 changes into intermediate-pressure refrigerant through isenthalpic decompression of the first expansion valve 13 which is in the decompression state. The intermediate-pressure refrigerant flowing out of the first expansion valve 13 flows through in an order: the gas-liquid separator 14→the second open-close valve 16 b, and then the refrigerant flows into the exterior heat exchanger 20. Here, the intermediate pressure is set such that the temperature of the intermediate-pressure refrigerant becomes lower than the temperature of the outside air.

The intermediate-pressure refrigerant flowing into the exterior heat exchanger 20 absorbs heat and evaporates by heat exchange with outside air blown by the blower fan 21. The refrigerant flowing out of the exterior heat exchanger 20 is decompressed by the second expansion valve 22 without change in enthalpy of the refrigerant, and then flows into the interior evaporator 23. A subsequent refrigerant flow and corresponding state change of the refrigerant are similar to those of the cooling mode.

As described above, in the third dehumidifying-heating mode, air having been cooled and dehumidified in the interior evaporator 23 can be heated in the interior condenser 12 and can be blown into the vehicle compartment, similarly to the first and second dehumidifying-heating modes. Accordingly, dehumidifying and heating of the vehicle compartment can be performed.

Because the exterior heat exchanger 20 is used as an evaporator by reducing the open degree of the first expansion valve 13 in the third dehumidifying-heating mode, a heat absorption amount of refrigerant from outside air can be increased. Thus, a heat radiation amount of refrigerant in the interior condenser 12 can be increased, and the air heating capacity of the interior condenser 12 can be thereby improved relative to the second dehumidifying-heating mode.

(b)(iv) Fourth Dehumidifying-Heating Mode

When the target outlet temperature TAO becomes higher than a third reference temperature during the operation of the third dehumidifying-heating mode, a fourth dehumidifying-heating mode is performed. The fourth dehumidifying-heating mode is another example of the dehumidifying-heating mode. In the fourth dehumidifying-heating mode, the open degree of the first expansion valve 13 is adjusted to be smaller than in the third dehumidifying-heating mode, and the open degree of the second expansion valve 22 is fully open.

Similarly to the first to third dehumidifying-heating modes, high-pressure refrigerant discharged from the discharge port 11 c of the compressor 11 flows into the interior condenser 12 and radiates heat by heat exchange with air having been cooled and dehumidified in the interior evaporator 23. Accordingly, air to be blown into the vehicle compartment is heated in the interior condenser 12.

The refrigerant flowing out of the interior condenser 12 changes into low-pressure refrigerant having a temperature lower than an outside air temperature through isenthalpic decompression in the first expansion valve 13 which is in the decompression state. The low-pressure refrigerant flowing out of the first expansion valve 13 flows through in the order: the gas-liquid separator 14→the second open-close valve 16 b, and then the refrigerant flows into the exterior heat exchanger 20.

The low-pressure refrigerant flowing into the exterior heat exchanger 20 absorbs heat and evaporates by heat exchange with outside air blown by the blower fan 21. The refrigerant flowing out of the exterior heat exchanger 20 flows into the interior evaporator 23 without decompression because the second expansion valve 22 is fully open. A subsequent refrigerant flow and corresponding state change of the refrigerant are similar to those of the cooling mode.

As described above, in the fourth dehumidifying-heating mode, air having been cooled and dehumidified in the interior evaporator 23 can be heated in the interior condenser 12 and can be blown into the vehicle compartment, similarly to the first to third dehumidifying-heating modes. Accordingly, dehumidifying and heating of the vehicle compartment can be performed.

In the fourth dehumidifying-heating mode, similarly to the third dehumidifying-heating mode, the exterior heat exchanger 20 is used as an evaporator, and the open degree of the first expansion valve 13 is smaller than in the third dehumidifying-heating mode. Hence, a refrigerant evaporation temperature in the exterior heat exchanger 20 can be reduced.

Therefore, the temperature difference between refrigerant and outside air in the exterior heat exchanger 20 can be increased, and the heat absorption amount of refrigerant from outside air can be increased, relative to the third dehumidifying-heating mode. As a result, a, heat radiation amount of refrigerant in the interior condenser 12 can be increased, and the air heating capacity of the interior condenser 12 can be thereby improved relative to the third dehumidifying-heating mode.

(c) Heating Mode

Next, details of the heating mode performed at step S9 will be described with reference to FIGS. 5 to 9. At step S91 shown in FIG. 5, the air conditioning controller 40 determines control states of, for example, the expansion valves 13, 22, the air mix door 34 and the refrigerant circuit switching portion (16 a, 16 b, 16 c).

Specifically, the open degree of the first expansion valve 13 is reduced to decompress refrigerant, and the second expansion valve 22 is fully closed. The control state of the servomotor for the air mix door 34 is determined such that the air mix door 34 is actuated to close the bypass air passage 35 as shown in FIG. 2. Additionally, the first and third open-close valves 16 a, 16 c are open, and the second open-close valve 16 b is closed.

Accordingly, at step S12 shown in FIG. 4, when the air conditioning controller 40 outputs control signals and control voltages to the controlled components, a refrigerant circuit of the heat pump cycle 10 shown by solid arrows in FIG. 2 is provided.

Specifically, the heat pump cycle 10 is switched to a gas-injection cycle (economizer refrigerant cycle). In this cycle, refrigerant is compressed in stages by the two compression mechanisms which are the lower-stage compression mechanism and the higher-stage compression mechanism of the compressor 11. Moreover, intermediate-pressure refrigerant in the heat pump cycle 10 is combined with refrigerant discharged from the lower-stage compression mechanism, and the combined refrigerant is drawn into the higher-stage compression mechanism.

At step S92, the air conditioning controller 40 determines the target pressure TPd of the higher-pressure side refrigerant pressure Pd, and then performs a control operation of step S93. Here, the higher-pressure side refrigerant pressure Pd is a pressure between the discharge port 11 c of the compressor 11 and the inlet side of the first expansion valve 13. The target pressure TPd is determined by using a control map stored in the air conditioning controller 40 based on the target outlet temperature TAO determined at step S4 shown in FIG. 4, such that the temperature of air to be blown into the vehicle compartment becomes the target outlet temperature TAO

At step S93, the air conditioning controller 40 determines whether a present rotation rate Nc of the compressor 11 reaches a largest rotation rate Ncmax predetermined based on durability of the compressor 11. In other words, the air conditioning controller 40 determines whether the present rotation rate Nc is equal to the largest rotation rate Ncmax (Nc=Ncmax). When the present rotation rate Nc is not equal to the largest rotation rate Ncmax at step S93, the air conditioning controller 40 performs a control process of step S94 to perform a subcool control. When the present rotation rate Nc is equal to the largest rotation rate Ncmax (Nc=Ncmax) at step S93, a control operation of step S95 is performed.

The subcool control performed at step S94 will be described referring to the flowchart of FIG. 6. The subcool control is performed when the present rotation rate Nc is not equal to the largest rotation rate Ncmax at step S93. In other words, the subcool control is performed when the refrigerant discharge capacity of the compressor 11 can be increased more than a present refrigerant discharge capacity of the compressor 11.

At step S941 shown in FIG. 6, the air conditioning controller 40 determines a target supercooling degree TSC of refrigerant flowing out of the interior condenser 12, and then performs a control operation of step S942. Specifically, at step S941, the target supercooling degree TSC is determined based on a temperature and a pressure of the refrigerant flowing out of the interior condenser 12 such that the COP of the heat pump cycle 10 becomes largest.

At step S942, the air conditioning controller 40 determines whether the target supercooling degree TSC is higher than a present supercooling degree SC of the refrigerant flowing out of the interior condenser 12. Here, the present supercooling degree SC is calculated based, on a temperature and a pressure of the refrigerant flowing out of the interior condenser 12. When the present supercooling degree SC is lower than the target supercooling degree TSC at step S942, a control operation of step S944 is performed. When the present supercooling degree SC is not lower than the target supercooling degree TSC at step S942, a control operation of step S943 is performed.

Here, the supercooling degree SC of the present embodiment is defined as an absolute value of a difference between a present temperature of refrigerant in a liquid state and a saturation temperature of the refrigerant at a constant pressure. Hence, a temperature of liquid refrigerant decreases in accordance with increase of the supercooling degree SC. At step S943, the open degree of the first expansion valve 13 is increased by a predetermined degree from a present open degree of the first expansion valve 13, and then a control process of step S98 is performed. The increase of the open degree of the first expansion valve 13 causes a pressure of high-pressure side refrigerant to reduce, so that the supercooling degree SC of refrigerant flowing out of the interior condenser 12 reduces to approach the target supercooling degree TSC.

At step S944, the air conditioning controller 40 determines whether a present open degree of the first expansion valve 13 is larger than a smallest open degree of the first expansion valve 13. When the present open degree of the first expansion valve 13 is larger than the smallest open degree of the first expansion valve 13 at step S944, a control operation of step S945 is performed. At step S945, the open degree of the first expansion valve 13 is decreased by a predetermined degree from the present open degree of the first expansion valve 13, and then the control process of step S98 is performed. The decrease of the open degree of the first expansion valve 13 causes the pressure of high-pressure side refrigerant to increase, so that the supercooling degree SC of refrigerant flowing out of the interior condenser 12 increases to approach the target supercooling degree TSC.

When the present open degree of the first expansion valve 13 is not larger than the smallest open degree at step S944, i.e., when the present open degree of the first expansion valve 13 is equal to the smallest open degree at step S944, the present open degree of the first expansion valve 13 cannot be decreased. Therefore, the present open degree is kept, and the control process of step S98 is performed.

The subcool control is performed at step S94 when the refrigerant discharge capacity of the compressor 11 can be increased more than a present refrigerant discharge capacity thereof, and the open degree of the first expansion valve 13 is adjusted so that the supercooling degree SC approaches the target supercooling degree TSC. Accordingly, the COP approaches a largest value.

Next, at step S95 shown in FIG. 5, the air conditioning controller 40 determines whether a present open degree of the first expansion valve 13 is smaller than a largest open degree thereof. In other words, the air conditioning controller 40 determines whether the first expansion valve 13 is in the fully open state. When the present open degree of the first expansion valve 13 is smaller than the largest open degree at step S95, the air conditioning controller 40 performs a control process of step S96 to perform a quality control (dryness control). When the present open degree of the first expansion valve 13 is not smaller than the largest open degree at step S95, the air conditioning controller 40 performs a control process of step S97 to perform a PTC-heater control.

The quality control performed at step S96 will be described with reference to the flowchart of FIG. 7. The quality control is performed when refrigerant flowing out of the interior condenser 12 can be made to be in a gas-liquid two-phase state by increasing the present open degree of the first expansion valve 13. For example, the quality control is performed when the heating capacity of the interior condenser 12 is insufficient during the subcool control.

At step S961, the air conditioning controller 40 determines whether a present higher-pressure side refrigerant pressure Pd is smaller than the target pressure TPd determined at step S92. When the present higher-pressure side refrigerant pressure Pd is smaller than the target pressure TPd at step S961, the air conditioning controller 40 performs a control operation, of step S962. When the present higher-pressure side refrigerant pressure Pd is not smaller than the target pressure TPd at step S961, in other words, when the present higher-pressure side refrigerant pressure Pd is equal to or larger than the target pressure TPd, the air conditioning controller 40 performs a control operation of step S964.

At step S962, the air conditioning controller 40 determines whether a present open degree of the first expansion valve 13 is smaller than the largest open degree thereof. In other words, the air conditioning controller 40 determines whether the first expansion valve 13 is in the fully open state. When the present open degree of the first expansion valve 13 is smaller than the largest open degree at step S962, the air conditioning controller 40 performs a control operation of step S963. At step S963, the present open degree of the first expansion valve 13 is increased by a predetermined degree. Subsequently, the control process of step S98 is performed.

When the present open degree of the first expansion valve 13 is not smaller than the largest open degree at step S962, in other words, when the first expansion valve 13 is in the fully open state, the present open degree of the first expansion valve 13 cannot be increased. Thus, the present open degree is kept, and the control process of step S98 is performed.

At step S964, the air conditioning controller 40 determines whether a present open degree of the first expansion valve 13 is larger than the smallest open degree of the first expansion valve 13. When the present open degree of the first expansion valve 13 is larger than the smallest open degree at step S964, the air conditioning controller 40 performs a control operation of step S965. At step S965, the present open degree of the first expansion valve 13 is increased by a predetermined degree. Then, the air conditioning controller 40 performs the control process of step S98.

When the present open degree of the first expansion valve 13 is not larger than the smallest open degree at step S964, in other words, when the present open degree of the first expansion valve 13 is in a fully closed state, the present open degree of the first expansion valve 13 cannot be reduced. Hence, the present open degree is kept, and the control process of step S98 is performed.

The quality control is performed at step S96 when the refrigerant discharge capacity of the compressor 11 cannot be increased more than the present refrigerant discharge capacity thereof, and the open degree of the first expansion valve 13 is increased to increase a flow amount (gas injection amount) of refrigerant flowing into the compressor 11 via the intermediate pressure port 11 b. Accordingly, a compression work amount of the compressor 11 is increased, and a quality (dryness) of refrigerant flowing out of the interior condenser 12 is increased. As a result, air to be blown into the vehicle compartment is heated to the target outlet temperature TAO.

Next, the PTC-heater control (control of the air heating capacity of the PTC heater 50) performed at step S97 shown in FIG. 5 will be described with reference to the flowchart of FIG. 8. The PTC-heater control is performed when the air heating capacity of the interior condenser 12 cannot be increased by a heating capacity control of the interior condenser 12, such as the subcool control and the quality control, which is performed for increasing the air heating capacity of the interior condenser 12. Specifically, when the rotation rate Nc of the compressor 11 is equal to the largest rotation rate Ncmax, and when the open degree of the first expansion valve 13 is equal to the largest open degree, air to be blown into the vehicle compartment cannot be heated to the target outlet temperature TAO by the control of the rotation rate Nc of the compressor 11 and the control of the open degree of the first expansion valve 13. In this case, the PTC-heater control shown in FIG. 8 is performed.

At step S971, the air conditioning controller 40 determines whether a present higher-pressure side refrigerant pressure Pd is higher than the target pressure TPd determined at step S92. When the present higher-pressure side refrigerant pressure Pd is higher than the target pressure TPd at step S971, the air conditioning controller 40 performs a control operation of step S972. When the present higher-pressure side refrigerant pressure Pd is not higher than the target pressure TPd at step S971, the air conditioning controller 40 performs a control operation of step S975.

When the present higher-pressure side refrigerant pressure Pd is higher than the target pressure TPd at step S971, the blown air can be heated to the target temperature TAO only by the air heating capacity of the interior condenser 12. Thus, the air conditioning controller 40 determines the operation mode of the PTC heater 50 at step S972, and then reduces the air heating capacity of the PTC heater 50.

Specifically, when the operation mode of the PTC heater 50 is determined to be the above-described HIGH mode at step S972, the air conditioning controller 40 (heating capacity control portion 40 a) switches the operation mode to the above-described LOW mode at step S973, and then performs the control process of step S98. When the operation mode of the PTC heater 50 is determined to be the LOW mode at step S972, the air conditioning controller 40 (heating capacity control portion 40 a) switches the operation mode to the above-described OFF mode at step S974, and then performs the control process of step S98.

Additionally, when the operation mode of the PTC heater 50 is determined to be the OFF mode at step S972, the OFF mode is kept, and the control process of step S98 is performed.

At step S971, when the present higher-pressure side refrigerant pressure Pd is higher than the target pressure TPd, the blown air cannot be heated to the target outlet temperature TAO only by the air heating capacity of the interior condenser 12. Thus, the air conditioning controller 40 determines the operation mode of the PTC heater 50 at step S975, and then increases the air heating capacity of the PTC heater 50.

Specifically, when the operation mode of the PTC heater 50 is determined to be the HIGH mode at step S975, the HIGH mode is kept, and the control process of step S98 is performed. When the operation mode of the PTC heater 50 is determined to be the LOW mode at step S975, the air conditioning controller 40 (heating capacity control portion 40 a) switches the operation mode to the HIGH mode at step S976, and then performs the control process of step S98. Additionally, when the operation mode of the PTC heater 50 is determined to be the OFF mode at step S975, the air conditioning controller 40 (heating capacity control portion 40 a) switches the operation mode to the LOW mode at step S977, and then performs the control process of step S98.

At step S98 shown in FIG. 5, the rotation rate Nc of the compressor 11 is determined by the feedback control so that the higher-pressure side refrigerant pressure Pd approaches the target pressure TPd. The determination of the rotation rate Nc of the compressor 11 at step S98 will be described referring to the flowchart of FIG. 9. At step S981, the air conditioning controller 40 determines whether a present higher-pressure side refrigerant pressure Pd is lower than the target pressure TPd determined at step S92.

When the present higher-pressure side refrigerant pressure Pd is determined to be lower than the target pressure TPd at step S981, a control operation of step S982 is performed. At step S982, the air conditioning controller 40 determines whether a present rotation rate Nc of the compressor 11 is lower than the largest rotation rate Ncmax. When the present rotation rate Nc of the compressor 11 is determined to be lower than the largest rotation rate Ncmax at step S982, a control operation of step S983 is performed. The air conditioning controller 40 increases the rotation rate Nc of the compressor 11 by a predetermined degree at step S983, and then performs the control operation of step S10 shown in FIG. 4.

When the present rotation rate Nc of the compressor 11 is determined not to be lower than the largest rotation rate Ncmax at step S982, in other words, when the present rotation rate Nc of the compressor 11 is equal to the largest rotation rate Ncmax, the present rotation rate Nc of the compressor 11 cannot be increased. Hence, the present rotation rate Nc is kept, and the control operation of step S10 shown in FIG. 4 is performed.

When the present higher-pressure side refrigerant pressure Pd is determined not to be lower than the target pressure TPd at step S981, a control operation of step S984 is performed. At step S984, the air conditioning controller 40 decreases the rotation rate Nc of the compressor 11 by a predetermined degree, and then performs the control operation of the step S10 shown in FIG. 4.

In the heating mode, because the control process is performed as described above, a state of refrigerant in the heat pump cycle 10 is changed as shown by the Mollier diagram of FIG. 10. In FIG. 10, a state change of refrigerant in the subcool control is shown by a bold solid line, and a state change of refrigerant in the quality control is shown by a bold dash line. Furthermore, a state change of refrigerant in the PTC-heater control is shown by a bold alternate long and short dash line in FIG. 10.

When the subcool control is performed in the heating mode as shown in the control process of step S94 in FIG. 6, high-pressure refrigerant flowing out of the discharge port 11 c of the compressor 11, which is shown by the point a in FIG. 10, flows into the interior condenser 12. The refrigerant flowing into the interior condenser 12 radiates heat and condenses through heat exchange with air having passed through the interior evaporator 23, as shown by the point a→the point b in FIG. 10. Accordingly, the air to be blown into the vehicle compartment is heated.

The refrigerant flowing out of the interior condenser 12 flows into the first expansion valve 13 which is in the decompression state, and changes into intermediate-pressure refrigerant through isenthalpic expansion and decompression in the first expansion valve 13 as shown by the point b→the point c1. Subsequently, the intermediate-pressure refrigerant decompressed in the first expansion valve 13 is separated into liquid refrigerant and gas refrigerant in the gas-liquid separator 14 as shown by the point c1→the point c2, and the point c1→the point c3 in FIG. 10.

The intermediate-pressure gas refrigerant separated in the gas-liquid separator 14 flows into the intermediate pressure port 11 b of the compressor 11 via the intermediate pressure passage 15 as shown by the point c2→the point a2 in FIG. 10 because the first open-close valve 16 is open. The refrigerant flowing into the compressor 11 via the intermediate pressure port 11 b is combined with refrigerant (the point a1 in FIG. 10) discharged from the lower-stage compression mechanism. The combined refrigerant is drawn into the higher-stage compression mechanism.

The intermediate-pressure liquid refrigerant separated in the gas-liquid separator 14 flows into the fixed throttle 17 because the second open-close valve 16 b is closed. The liquid refrigerant changes into low-pressure refrigerant through isenthalpic expansion and decompression in the fixed throttle 17 as shown by the point c3→the point c4 in FIG. 10. The low-pressure refrigerant flowing out of the fixed throttle 17 flows into the exterior heat exchanger 20. Then, the low-pressure refrigerant absorbs heat and evaporates through heat exchange with outside air blown by the blower fan 21 in the exterior heat exchanger 20 as shown by the point c4→the point d in FIG. 10.

The refrigerant flowing out of the exterior heat exchanger 20 flows into the accumulator 24 via the bypass passage 25 because the third open-close valve 16 c is open. The refrigerant is separated into gas refrigerant and liquid refrigerant in the accumulator 24, and the separated gas refrigerant is drawn into the compressor 11 through the suction port 11 a, as shown by the point e in FIG. 10, to be compressed in the compressor 11. The separated liquid refrigerant is accumulated in the accumulator 24 as surplus refrigerant which is unnecessary refrigerant to provide a required refrigeration capacity of the heat pump cycle 10.

Here, the reason, why the point d and the point e are different from each other in FIG. 10, is that a pressure loss is generated in the gas refrigerant passing through a refrigerant pipe from the accumulator 24 to the suction port 11 a of the compressor 11. Ideally, the points d and e are identical with each other. The reason of the difference is similar to the other air conditioning modes.

Thus, in the subcool control of the heating mode, air can be heated through heat exchange with high-temperature and high-pressure refrigerant discharged from the compressor 11 in the interior condenser 12, and can be blown into the vehicle compartment. Accordingly, the vehicle compartment can be heated. Moreover, in the subcool control, the super cooling degree SC of refrigerant flowing out of the interior condenser 12, shown by the point b in FIG. 10, can be controlled to approach the target super cooling degree TSC through the adjustment of the open degree of the first expansion valve 13, and the COP of the heat pump cycle 10 can thereby approach a largest value.

In the subcool control, when the rotation rate Nc of the compressor 11 is equal to the largest rotation rate Ncmax, and when the air heating capacity of the interior condenser 12 is insufficient to increase a temperature of air blown to the vehicle compartment to the target outlet temperature TAO, the subcool control is switched to the quality control shown by the control process of step S96 in FIG. 7.

When the quality control is performed, a state of refrigerant is changed as shown by the bold dash line in FIG. 10. In FIG. 10, a state of refrigerant in the quality control is assigned the same character as a corresponding state in the subcool control, and the characters in the quality control are apostrophized.

In the quality control, because the open degree of the first expansion valve 13 is increased to increase a quality of refrigerant flowing out of the interior condenser 12, a state of refrigerant flowing out of the interior condenser 12 changes into a state shown by the point b′ in FIG. 10. Moreover, a pressure of refrigerant flowing into the compressor 11 via the intermediate pressure port 11 b and a pressure refrigerant discharged from the compressor 11 via the discharge port 11 c are increased as compared with the case of the subcool control, as shown by, for example, the points c2′ and a′ in FIG. 10.

Thus, in the quality control, a temperature of refrigerant discharged from the compressor 11 can be increased, and a temperature difference between high-pressure refrigerant flowing through the interior condenser 12 and air flowing into the interior condenser 12 can be thereby widen as compared with the case of the subcool control. Additionally, a flow amount (gas injection amount) of refrigerant flowing into the compressor 11 via the intermediate pressure port 11 b can be increased as compared with the case of the subcool control. As a result, in the quality control, the air heating capacity of the interior condenser 12 can be increased as compared with the case of the subcool control.

Here, as described above, the air heating capacity of the interior condenser 12 can be expected to be increased in the quality control. However, in the quality control, an enthalpy difference between refrigerant flowing at the refrigerant inlet of the interior condenser 12 and refrigerant flowing at the refrigerant outlet of the interior condenser 12 is reduced as compared with the case of the subcool control, and the air heating capacity of the interior condenser 12 may be thereby unable to be increased when the open degree of the first expansion valve 13 is larger than a certain value.

In the present embodiment, during the quality control, when the open degree of the first expansion valve 13 is equal to the largest open degree, and when air to be blown to the vehicle compartment cannot be heated to the target outlet temperature TAO by the air heating capacity of the interior condenser 12, the quality control is switched to the PTC-heater control shown by the control process of step S97 in FIG. 8. In other words, when a temperature of air that is to be blown into the vehicle compartment is equal to or lower than the target outlet temperature TAO in the quality control, the PTC-heater control is performed.

When the PTC-heater control is performed, a state of refrigerant is changed as shown by the bold alternate long and short dash line in FIG. 10. In FIG. 10, a state of refrigerant in the PTC-heater control is assigned the same character as a corresponding state in the subcool control, and the characters in the PTC-heater control are double-apostrophized.

In the PTC-heater control, the air conditioning controller 40 (the heating capacity control portion 40 a) increases a voltage applied to the PTC heater 50 to increase the air heating capacity of the PTC heater 50. Accordingly, a temperature of air flowing into the interior condenser 12 is increased, and a heat absorption amount of the air from the refrigerant in the interior condenser 12, i.e., a heat radiation amount of the refrigerant to the air in the interior condenser 12 reduces temporarily.

Thus, in the PTC-heater control, the heat exchange capacity of the interior condenser 12 decreases substantially, and refrigerant circulating in the heat pump cycle 10 balances so that a pressure of refrigerant in the interior condenser 12 increases, as shown by the point a″ and the point b″ in FIG. 10. Therefore, a temperature of refrigerant discharged from the compressor 11 can be increased, and a temperature difference between high-pressure refrigerant flowing through the interior condenser 12 and air flowing into the interior condenser 12 can be increased.

Additionally, a compression work amount in a compression process in the higher-stage compression mechanism of the compressor 11 can be increased. Here, the compression process in the higher-stage compression mechanism is a compression process from the intermediate pressure port 11 b to the discharge port 11 c and is shown by the point a2′→the point a″ in FIG. 10. Hence, the enthalpy difference between refrigerant in the refrigerant inlet of the interior condenser 12 and refrigerant in the refrigerant outlet of the interior condenser 12 can be increased as compared with the case of the quality control, as shown by the difference Δic2′→the difference Δic2″ in FIG. 10. As a result, in the PTC-heater control, the air heating capacity in the interior condenser 12 can be improved as compared with the case of the quality control.

The vehicle air conditioner 1 of the present embodiment, as described above, can provide cooling, dehumidifying-heating and heating of the vehicle compartment, and can heat air efficiently and effectively depending on a required air heating capacity in the dehumidifying-heating mode and the heating mode.

In the present embodiment, the PTC heater 50 used as an example of the auxiliary heater is arranged upstream of the interior condenser 12 in the air flow direction so as to heat air before the air is heated by the interior condenser 12. Thus, the air heating capacity in the interior condenser 12 can be improved in the PTC-heater control. In this case, power consumption in the PTC heater 50 can be reduced relative to a configuration in which the PTC heater 50 heats air having been heated in the interior condenser 12.

More specifically, if the PTC heater 50 is arranged to heat air having been heated in the interior condenser 12, the PTC heater 50 is required to generate a heat of 2kW (standard heating capacity) when a rated voltage is applied. However, in the present embodiment, the PTC heater 50 may be configured to generate a heat of 800W when the rated voltage is applied. Therefore, the power consumption in the PTC heater 50 can be reduced relatively.

Furthermore, because a PTC heater having an air heating capacity lower than the standard heating capacity can be adopted as the PTC heater 50, the vehicle air conditioner 1 (refrigerant cycle device) can be reduced in size and in manufacturing cost as a whole by downsizing the PTC heater 50 and by reducing a diameter of harness (electric power line) connecting the PTC heater 50 and the air conditioning controller 40, for example.

In the present embodiment, as in the description of control process of step S97 shown in FIG. 8, the heating capacity control portion 40 a adjusts the air heating capacity of the PTC heater 50 in the PTC-heater control so that the higher-pressure side refrigerant pressure Pd (refrigerant pressure in the interior condenser 12) becomes equal to the target pressure TPd, and the target pressure TPd is determined based on the target outlet temperature TAO. Therefore, air to be blown into the vehicle compartment can be heated to the target outlet temperature TAO easily, and unnecessary energy consumption can be limited.

In the present embodiment, the cycle configuration of the heat pump cycle 10 is switchable variedly depending on the air conditioning mode, and specifically can provide the gas injection cycle at least in the heating mode. When the gas injection cycle is provided at least in the heating mode of the heat pump cycle 10 of the present embodiment, the air heating capacity of the interior condenser 12 in the heating mode of the present embodiment can be improved reliably.

This will be described by comparing the vehicle air conditioner 1 of the present embodiment with a vehicle air conditioner of a comparative example. The vehicle air conditioner of the comparative example includes a normal vapor-compressing refrigerant cycle having a compressor, a radiator corresponding to the interior condenser 12 of the present embodiment, an expansion valve, and an evaporator corresponding to the exterior heat exchanger 20 of the present embodiment. These components of the refrigerant cycle are connected in a loop shape. The vehicle air conditioner of the comparative example further includes a PTC heater arranged upstream of the radiator to heat air that is to flow into the radiator.

In FIG. 11, a bold solid line shows a state change of refrigerant in a case where a supercooling degree of refrigerant flowing out of the radiator is controlled to approach a target supercooling degree without supply of electric power to the PTC heater. A bold dash line shows a state change of refrigerant in a case where the supercooling degree of refrigerant flowing out of the radiator is controlled to approach the target supercooling degree with the supply of electric power to the PTC heater. In FIG. 11, a state of refrigerant is assigned the same character as the corresponding state of refrigerant in FIG. 10.

By the supply of electric power to the PTC heater, the higher-pressure side refrigerant pressure Pd is increased, and a compression work amount in the compressor can be increased as shown by the difference Δic→the difference Aid in FIG. 11. However, an enthalpy difference (heat absorption amount of the exterior heat exchanger 20) between refrigerant at a refrigerant inlet of the exterior heat exchanger 20 and refrigerant at a refrigerant outlet of the exterior heat exchanger 20 is reduced as shown by the difference Δie→the difference Δie′ in FIG. 11. Consequently, an air heating capacity of the radiator may decrease in the comparative example.

In contrast, in the present embodiment, because the gas injection cycle is provided at least in the heating mode, the air heating capacity in the interior condenser 12 can be improved certainly by arranging the auxiliary heater (e.g., PTC heater 50) such that air is heated by the auxiliary heater before being heated through heat exchange with high-pressure refrigerant in the interior condenser 12.

Second Embodiment

In the above-described first embodiment, the PTC heater 50 is adopted as an example of the auxiliary heater. In a second embodiment, as shown in FIG. 12, a vehicle air conditioner 1 includes an auxiliary heat exchanger 60 as an example of the auxiliary heater, instead of the PTC heater 50 of the first embodiment. The auxiliary heat exchanger 60 heats air by using a coolant (heat medium) as a heat source. The coolant cools a non-shown electric motor for vehicle running, and cools a non-shown inverter which supplies electric power to the vehicle-running electric motor. Thus, the vehicle-running electric motor and the inverter are used as examples of an external heat source.

In FIG. 12, the refrigerant cycle of the heat pump cycle 10 is set in a state for the heating mode. In FIG. 12, a part is assigned the same numeral as a same or equivalent part of the first embodiment.

The auxiliary heat exchanger 60 is arranged in a coolant circuit 61 through which the coolant circulates to cool the external heat source such as the vehicle-running electric motor and the inverter. The auxiliary heat exchanger 60 is a tank-and-tube type heat exchanger which heats air through heat exchange with the coolant flowing in the auxiliary heat exchanger 60. The auxiliary heat exchanger 60 is arranged upstream of the interior condenser 12 to heat air before flowing into the interior condenser 12.

Moreover, the auxiliary heat exchanger 60 has an air heating capacity lower than a standard heating capacity, similarly to the PTC heater 50 of the first embodiment. The standard heating capacity of the present embodiment is defined as a necessary heating capacity (largest heating capacity) for heating air to the target outlet temperature TAO by using both the air heating capacity of the interior condenser 12 and the air heating capacity of the auxiliary heat exchanger 60 in a case where the auxiliary heat exchanger 60 heats air which has been heated in the interior condenser 12.

In the present embodiment, the auxiliary heat exchanger 60 includes a heat-exchange core portion in which the coolant exchanges heat with air, and the heat-exchange core portion has a heat-exchange area smaller than a necessary heat-exchange area for providing the auxiliary heat exchanger 60 with the standard heating capacity. Consequently, the auxiliary heat exchanger 60 has the air heating capacity lower than the standard heating capacity.

In the coolant circuit 61, a flow control valve 62 is provided to adjust a flow amount of the coolant flowing into the auxiliary heat exchanger 60. The flow control valve 62 is an electric open-degree control valve which includes a valve body and an electric actuator being able to actuate the valve body to change a cross section of a coolant passage of the coolant circuit 61. An operation of the flow control valve 62 is controlled by a control signal outputted from the air conditioning controller 40.

Thus, the air conditioning controller 40 controls the operation of the flow control valve 62, and a flow amount of the coolant flowing into the auxiliary heat exchange 60 is thereby adjusted. Accordingly, the air heating capacity of the auxiliary heat exchanger 60 is adjusted. Therefore, the flow control valve 62 of the present embodiment is used as an example of the heating capacity adjusting portion.

The air conditioning controller 40 of the present embodiment is capable of switching an operation mode of the flow control valve 62. The operation mode of the flow control valve 62 includes a HIGH mode, a LOW mode and an OFF mode. In the HIGH mode, the flow control valve 62 fully opens the coolant passage of the coolant circuit 61 to set the air heating capacity of the auxiliary heat exchanger 60 relatively high. In the LOW mode, the flow control valve 62 moderately opens the coolant passage to set the air heating capacity of the auxiliary heat exchange 60 relatively low. In the OFF mode, the flow control valve 62 closes the coolant passage so that the auxiliary heat exchanger 60 is not provided with an air heating capacity.

The operation mode of the flow control valve 62 is switched at step S97 shown in FIG. 5, similarly to the PTC heater 50 of the first embodiment. The other configurations and operations of the vehicle air conditioner 1 of the second embodiment are similar to that of the vehicle air conditioner 1 of the first embodiment.

Accordingly, the vehicle air conditioner 1 (refrigerant cycle device) is capable of improving the air heating capacity in the interior condenser 12, similarly to the first embodiment. As a result, an energy amount consumed by the auxiliary heater (60) in the heating mode can be reduced, and a size and a manufacturing cost of the vehicle air conditioner 1 can be reduced as a whole.

Third Embodiment

In the first embodiment, the air conditioning controller 40 controls an electric power (e.g. electric voltage) supplied to the PTC heater 50, thereby switching the operation mode of the PTC heater 50 among the HIGH mode, the LOW mode and the OFF mode. In a third embodiment, multiple PTC heaters (electric heaters) are integrated into a single PTC heater 50.

Specifically, three PTC heaters are integrated into the PTC heater 50, and the air conditioning controller 40 changes the number of PTC heaters, which are energized, to control an air heating capacity of the PTC heater 50.

In other words, by changing the energized number of PTC heaters, an electric power amount supplied to the PTC heater 50 is adjusted. The PTC heater 50 of the present embodiment has an air heating capacity lower than the standard heating capacity even when all of the three PTC heaters are energized.

The energized number of PTC heaters is determined at step S97 shown in FIG. 13. The control process of step S97 in FIG. 13 corresponds to the control process of step S97 in FIG. 8 described in the first embodiment. When a present higher-pressure side refrigerant pressure Pd is determined to be lower than the target pressure TPd at step S971, a control operation of step S972′ is performed to increase the energized number of PTC heaters. For example, when the energized number of PTC heaters is three at step S972′ in a case where the total number of the PTC heaters is three, the energized number of PTC heaters is kept three.

When the present higher-pressure side refrigerant pressure Pd is determined not to be lower than the target pressure TPd at step S971, a control operation of step S975′ is performed to decrease the energized number of PTC heaters. Here, when the energized number of PTC heaters is zero at step S975′, in other words, when the PTC heater 50 is not energized, the PTC heater 50 is kept in a non-energized state.

The other configurations and the other operations in the third embodiment are similar to those of the first embodiment. Thus, in a vehicle air conditioner (refrigerant cycle device) of the third embodiment, similar effects to the first embodiment can be obtained. Furthermore, the air heating capacity of the PTC heater 50 can be changed in stages (e.g., three stages) by changing the energized number of PTC heaters, and an energy amount consumed by the auxiliary heater (PTC heater 50) in the heating mode can be thereby reduced further.

Fourth Embodiment

In the above-described first to third embodiments, the heat pump cycle 10 is a two stage expansion-type gas injection cycle, and includes the first expansion valve 13 as an example of the higher-pressure side expansion device, the fixed throttle 17 as an example of the lower-pressure side expansion device, and the gas-liquid separator 14. In the first and third embodiments, the gas-liquid separator 14 separates intermediate-pressure refrigerant, which has been decompressed by the first expansion valve 13, into gas refrigerant and liquid refrigerant, and the separated gas refrigerant flows to the intermediate pressure port 11 b. In a fourth embodiment, a heat pump cycle 10 is an inner heat-exchange gas injection cycle, and does not include the first expansion valve 13, the fixed throttle 17 and the gas-liquid separator 14.

A vehicle air conditioner 1 (refrigerant cycle device) of the fourth embodiment will be described with reference to FIG. 14. In the fourth embodiment, the heat pump cycle 10 of the vehicle air conditioner 1 includes a refrigerant branch portion 70 provided in a refrigerant passage connected to the refrigerant outlet side of the interior condenser 12. High-pressure refrigerant flowing out of the interior condenser 12 passes through the refrigerant passage, and the refrigerant passage branches into multiple passages in the refrigerant branch portion 70. In the fourth embodiment, the refrigerant passage branches into first and second refrigerant passages 71 and 73 at the refrigerant branch portion 70. The heat pump cycle 10 further includes a thermostatic expansion valve 72 as an example of a first expansion device that is provided in the first refrigerant passage 71 and decompresses high-pressure refrigerant flowing out of the interior condenser 12 so that the high-pressure refrigerant changes into intermediate-pressure refrigerant.

The heat pump cycle 10 further includes an inner heat exchanger 74 in which high-pressure refrigerant flowing through the second refrigerant passage 73 exchanges heat with the intermediate-pressure refrigerant decompressed by the thermostatic expansion valve 72. The inner heat exchanger 74 includes a high-pressure passage portion 74 a, through which the high-pressure refrigerant from the second refrigerant passage 73 flows, and an intermediate-pressure passage portion 74 b, through which the intermediate-pressure refrigerant decompressed by the thermostatic expansion valve 72 flows.

In the inner heat exchanger 74, the high-pressure refrigerant having a relatively high temperature and flowing in the high-pressure passage portion 74 a heats the intermediate-pressure refrigerant having a relatively low temperature and flowing in the intermediate-pressure passage portion 74 b. Accordingly, the intermediate-pressure refrigerant is evaporated to be gas refrigerant.

A refrigerant outlet side of the intermediate-pressure passage portion 74 b is connected to the intermediate pressure port 11 b of the compressor 11 through an intermediate pressure passage 15. A thermostatic portion 72 a of the thermostatic expansion valve 72 is provided in or adjacent to the intermediate pressure passage 15. The thermostatic expansion valve 72 has a valve body that moves due to a pressure of intermediate-pressure refrigerant passing through the thermostatic expansion valve 72 and due to a pressure depending on a temperature of intermediate-pressure refrigerant detected by the thermostatic portion 72 a. Accordingly, an open degree of the thermostatic expansion valve 72 is automatically adjusted so that the intermediate-pressure refrigerant flowing out of the intermediate-pressure passage portion 74 b has a predetermined superheat degree.

The intermediate-pressure gas refrigerant, which has evaporated in the intermediate-pressure passage portion 74 b of the inner heat exchanger 74, passes through the intermediate pressure passage 15, and then flows into the compressor 11 via the intermediate pressure port 11 b.

A refrigerant outlet side of the high-pressure passage portion 74 a is connected to an inlet side of an electric expansion valve 75. The electric expansion valve 75 is used as an example of a second expansion device that decompresses the high-pressure refrigerant flowing out of the high-pressure passage portion 74 a into low-pressure refrigerant, and an open degree of the electric expansion valve 75 is adjustable electrically. An outlet side of the electric expansion valve 75 is connected to the refrigerant inlet side of the exterior heat exchanger 20. The electric expansion valve 75 may have a similar configuration to that of the first expansion valve 13 described in the first embodiment.

The other configurations of the vehicle air conditioner 1 of the fourth embodiment are similar to those of the first embodiment. Thus, in FIG. 14, parts same as or similar to parts of the first embodiment are assigned the same numerals as the parts of the first embodiment, and descriptions of the parts are omitted in the fourth embodiment.

In FIG. 14, a refrigerant temperature sensor 41 a and a refrigerant pressure sensor 41 b of an air conditioning control sensor group 41 are provided in the refrigerant passage connected to the refrigerant outlet side of the interior condenser 12. The refrigerant temperature sensor 41 a detects a temperature of high-pressure refrigerant flowing out of the interior condenser 12, and the refrigerant pressure sensor 41 b detects a pressure of high-pressure refrigerant flowing out of the interior condenser 12.

In the fourth embodiment, the air conditioning controller 40 determines a state of high-pressure refrigerant flowing out of the interior condenser 12 (i.e., a supercooling degree or a quality (dryness) of the high-pressure refrigerant) based on detection signals from the sensors 41 a and 41 b. The open degree of the electric expansion valve 75 is adjusted depending on the state of the high-pressure refrigerant.

In the heating mode of the inner heat-exchange gas injection cycle of the fourth embodiment, the subcool control or the quality control (dryness control) shown in FIGS. 5 to 7 is performed by controlling the open degree of the electric expansion valve 75 until the rotation rate Nc of the compressor 11 becomes equal to the largest rotation rate Ncmax and until the open degree of the electric expansion valve 75 becomes largest. When the air conditioning controller determines that the heating capacity of the interior condenser 12 is insufficient in the state where the rotation rate Nc is equal to the largest rotation rate Ncmax and where the open degree of the electric expansion valve 75 is largest, the PTC heater 50 is activated. In other words, when the heating capacity of the interior condenser 12 cannot be increased by the heating capacity control of the interior condenser 12, such as the subcool control and the quality control, the PTC heater 50 is activated.

When the PTC heater 50 is operated in the heating mode, air that is to be blown into the vehicle compartment is heated firstly by the PTC heater 50, and is heated subsequently by the interior condenser 12 in the air conditioning unit 30.

When the PTC heater 50 is operated in the heating mode, a temperature of air flowing into the interior condenser 12 is higher than that in a case where the PTC heater 50 is not in operation. Hence, refrigerant cycle in the heat pump cycle 10 is balanced so that a pressure of high-pressure refrigerant discharged from the compressor 11 and a condensation temperature of the refrigerant become higher than those in the case where the PTC heater 50 is not in operation. Because the pressure of the high-pressure refrigerant is increased, the compression work amount of the compressor 11 is increased. As a result, the air heating capacity in the interior condenser 12 can be improved.

FIG. 15 shows refrigerant states in the heat pump cycle 10 of the fourth embodiment during the heating mode. In FIG. 15, a solid line shows refrigerant states when the PTC heater 50 is not operated, and an alternate long and short dash line shows refrigerant states when the PTC heater 50 is operated.

Some of reference characters in FIG. 15 are same as those in FIG. 10. Each of the same reference characters between FIGS. 15 and 10 indicates a state of refrigerant flowing in the same component of the heat pump cycle 10. Reference characters f and f″ indicate states of refrigerant flowing in the refrigerant branch portion 70. A reference character g indicate a state of refrigerant flowing in an outlet portion of the thermostatic expansion valve 72, i.e., a state of refrigerant flowing in an inlet portion of the intermediate-pressure passage portion 74 b of the inner heat exchanger 74.

Reference characters h and h″ indicate states of refrigerant flowing in an outlet portion of the high-pressure passage portion 74 a of the inner heat exchanger 74, i.e. states of refrigerant flowing in an inlet portion of the electric expansion valve 75. A reference character c indicates a state of refrigerant flowing in an outlet portion of the electric expansion valve 75, i.e., a state of refrigerant flowing in an inlet portion of the exterior heat exchanger 20. An enthalpy increment indicated by a line between points g and a2 in FIG. 15 and enthalpy decrements indicated by lines between points f and h and between points f″ and h″ in FIG. 15 are based on inner heat exchange in the inner heat exchanger 74.

In the fourth embodiment, similarly to the above-described embodiments, when the PTC heater 50 is operated, the compression work amount (Δic2″ in FIG. 15) in the higher-stage compression mechanism of the compressor 11 can be made to be larger than that (Δic2 in FIG. 15) in the case where the PTC heater 50 is not in operation. Consequently, the air heating capacity in the interior condenser 12 can be improved.

In the fourth embodiment, an electric heater such as the PTC heater 50 is used as an example of the auxiliary heater. Alternatively, the auxiliary heat exchanger 60 described in the second embodiment and shown in FIG. 12 may be used as an example of the auxiliary heater in the fourth embodiment.

Fifth Embodiment

In the first to fourth embodiments, the PTC heater 50 or the auxiliary heat exchanger 60, which is an example of the auxiliary heater, is arranged upstream of the interior condenser 12 in the air flow direction. In a fifth embodiment, a PTC heater 50 that is an example of the auxiliary heater is arranged in parallel with an interior condenser 12 in the air flow direction. In other words, the PTC heater 50 and the interior condenser 12 are arranged in a direction perpendicular to the air flow direction.

Specifically, as shown in FIG. 16, a heat exchange portion of the interior condenser 12, through which high-pressure refrigerant flows, is separated into multiple heat exchange portions 12 a, 12 b and 12 c in the direction perpendicular to the air flow direction. Two PTC heaters 50 are arranged between the heat exchange portions 12 a and 12 b and between the heat exchange portions 12 b and 12 c respectively. Accordingly, as shown in FIG. 16, the multiple heat exchange portions 12 a, 12 b and 12 c and the two PTC heaters 50 are arranged in parallel with respect to the air flow direction in the fifth embodiment.

Each of the multiple heat exchange portions 12 a, 12 b and 12 c includes tubes in which high-pressure refrigerant flows, and fins arranged between the tubes. The multiple heat exchange portions 12 a, 12 b and 12 c and the two PTC heaters 50 are integrated with each other by arbitrary fixing method or the like in a state where the two PTC heaters 50 arranged between the multiple heat exchange portions 12 a, 12 b and 12 c. Therefore, the interior condenser 12 of the fifth embodiment is integrated with the PTC heaters 50.

Because the multiple heat exchange portions 12 a, 12 b and 12 c and the two PTC heaters 50 are arranged in parallel, with respect to the air flow direction in the interior condenser 12 of the fifth embodiment, the multiple heat exchange portions 12 a, 12 b and 12 c and the PTC heaters 50 simultaneously heat air flowing therethrough when the PTC heaters 50 are operated.

The interior condenser 12 has the above-described integrated structure. Additionally, the multiple heat exchange portions 12 a, 12 b and 12 c, through which high-pressure refrigerant flows, and the two PTC heaters 50 are arranged alternately in the direction perpendicular to the air flow direction. Therefore, a temperature of air flowing into the heat exchange portions 12 a, 12 b and 12 c is increased by heating effects of the PTC heaters 50 so as to be higher than that in a case where the PTC heaters 50 are not provided. Accordingly, a pressure of the high-pressure refrigerant and a condensation temperature of refrigerant can be increased, and a compression work amount of the compressor 11 can be increased. As a result, the air heating capacity in the interior condenser 12 can be improved.

In the fifth embodiment, an electric heater such as the PTC heater 50 is used as an example of the auxiliary heater. Alternatively, the auxiliary heat exchanger 60 described in the second embodiment and shown in FIG. 12 may be used as an example of the auxiliary heater in the fifth embodiment. In other words, the interior condenser 12 may have a heat exchanger configuration, such that the multiple heat exchange portions 12 a, 12 b and 12 c, through which high-pressure refrigerant flows, and the auxiliary heat exchanger 60, through which heat medium for cooling of an exterior heat source flows, may be arranged alternately and integrated with each other.

Although the present disclosure has been fully described in connection with the preferred embodiments thereof with reference to the accompanying drawings, it is to be noted that various changes and modifications will become apparent to those skilled in the art. Thus, the present disclosure is not limited to the above-described embodiments, and can be modified variedly as follows without departing from the scope of the present disclosure.

(1) In the above-described embodiments, the refrigerant cycle device of the present disclosure is used for the vehicle air conditioner 1, which is used for an electric vehicle. The refrigerant cycle device of the present disclosure can be suitably used for a vehicle, in which waste heat from an internal combustion engine is insufficient for use as a heat source of air heating. For example, the refrigerant cycle device of the present disclosure can be suitably used for a hybrid vehicle in which a driving force for vehicle running is obtained from the engine and an electric motor for vehicle running.

Moreover, the refrigerant cycle device of the present embodiment may be used for a stationary air conditioner, a cool temperature storage and a liquid heater, for example. When the refrigerant cycle device is used for the liquid heater, a liquid-refrigerant heat exchanger may be adopted as the above-described using-side heat exchanger, and a liquid pump or a flow control valve may be adopted as a flow controller which adjusts a flow amount of liquid flowing into the liquid-refrigerant heat exchanger.

(2) In the above-described second embodiment, the vehicle-running electric motor and the inverter are adopted as the external heat source, and the coolant (heat medium), which cools the vehicle-running electric motor and the inverter, is used as the heat source of the auxiliary heat exchanger 60. However, the external heat source and the heat medium are not limited to this. For example, when the refrigerant cycle device of the present disclosure is used for a vehicle air conditioner of the above-described hybrid vehicle, the engine of the hybrid vehicle may be adopted as the external heat source, and an engine coolant for the hybrid vehicle may used as the heat medium.

Further, when the refrigerant cycle device of the present disclosure is used for a stationary device, such as the stationary air conditioner, the cool temperature storage or the liquid heater, an engine for a compressor of the stationary device may be adopted as the external heat source, and other heat source of the stationary device may be adopted as the external heat source.

(3) In the above-described first and second embodiments, the air heating capacity of the auxiliary heater (e.g., the PTC heater 50, the auxiliary, heat exchanger 60) is changed in stages by switching the operation mode of the auxiliary heater among the HIGH mode, the LOW mode and the OFF mode. In the above-described third embodiment, the air heating capacity of the auxiliary heater is changed in multiple stages. However, the adjustment of the air heating capacity of the auxiliary heater is not limited to these. For example, the air heating capacity of the auxiliary heater may be increased gradually and continuously in accordance with increase of a value obtained by subtracting the higher-pressure side refrigerant pressure Pd from the target pressure TPd.

(4) In the above-described second embodiment, the auxiliary heat exchanger 60 has the heat-exchange area smaller than the necessary heat exchange area for providing the auxiliary heat exchanger 60 with the standard heating capacity, and the auxiliary heat exchanger 60 thus has the air heating capacity lower than the standard heating capacity. However, the auxiliary heat exchanger 60 is not limited to this. For example, the number of tubes of the auxiliary heat exchanger 60 (tank-and-tube type heat exchanger) or the number of fins for promotion of heat exchange may be reduced, or an efficiency of heat exchange may be reduced, so that the auxiliary heat exchanger 60 has the air heating capacity lower than the standard heating capacity. Moreover, another type of heat exchanger may be adopted as the auxiliary heat exchanger 60 alternatively.

(5) In the above-described embodiments, at step S6 in FIG. 4, the air conditioning mode is determined from among the cooling mode, the dehumidifying-heating mode and the heating mode depending on a state of the mode selecting switch. However, the determination of the air conditioning mode is not limited to this. For example, the cooling mode may be selected when the outside temperature is lower than the preset temperature, and the heating mode may be selected when the outside temperature is higher than the preset temperature.

(6) In the above-described embodiments, the quality X of refrigerant flowing into the exterior heat exchanger 20 is set to be equal to or lower than 0.1 in the heating mode by appropriately adjusting a flow characteristic of the fixed throttle 17 used as an example of the lower-pressure side expansion device. However, the lower-pressure side expansion device is not limited to the fixed throttle 17.

For example, a variable throttle mechanism having a similar structure to the first expansion valve 13 may be adopted as the lower-pressure side expansion device. In this case, the air conditioning controller 40 may detect the quality X of refrigerant flowing into the exterior heat exchanger 20 based on a temperature and a pressure of the refrigerant flowing into the exterior heat exchanger 20, and may control an open degree of the variable throttle mechanism as an example of the lower-pressure side expansion device so that the detected quality X becomes equal to or lower than 0.1.

(7) In the above-described embodiments, the dehumidifying-heating mode is switched from the first dehumidifying-heating mode to the fourth dehumidifying-heating mode in stages in accordance with increase of the target outlet temperature TAO, but the switching of the dehumidifying-heating mode is not limited to this. For example, the dehumidifying-heating mode may be switched from the first dehumidifying-heating mode to the fourth dehumidifying-heating mode in a continuous manner in accordance with increase of the target outlet temperature TAO.

Hence, the open degree of the first expansion valve 13 may be decreased, and the open degree of the second expansion valve 22 may be increased, in accordance with the increase of the target outlet temperature TAO. By changing the open degrees of both the first and second expansion valves 13, 22, a pressure and a temperature of refrigerant in the exterior heat exchanger 20 can be adjusted. Therefore, the exterior heat exchanger 20 can be switched from a state as a radiator to a state as an evaporator automatically.

Additional advantages and modifications will readily occur to those skilled in the art. The disclosure in its broader terms is therefore not limited to the specific details, representative apparatus, and illustrative examples shown and described. 

1. A refrigerant cycle device comprising: a compressor having a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion; a using-side heat exchanger which heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor; an intermediate pressure passage through which intermediate-pressure gas refrigerant obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger is introduced into the intermediate pressure port; an exterior heat exchanger in which low-pressure refrigerant obtained by decompression of the high-pressure refrigerant flowing out of the using-side heat exchanger evaporates, the exterior heat exchanger causing the evaporated low-pressure refrigerant to flow to the suction port; and an auxiliary heater which heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.
 2. A refrigerant cycle device comprising: a compressor having a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion; a using-side heat exchanger which heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor; a higher-pressure side expansion device configured to decompress the high-pressure refrigerant flowing out of the using-side heat exchanger into intermediate-pressure refrigerant; a gas-liquid separation portion configured to separate the intermediate-pressure refrigerant flowing out of the higher-pressure side expansion device into intermediate-pressure gas refrigerant and intermediate-pressure liquid refrigerant, the gas-liquid separation portion causing the separated intermediate-pressure gas refrigerant to flow to the intermediate pressure port; a lower-pressure side expansion device configured to decompress the separated intermediate-pressure liquid refrigerant flowing out of the gas-liquid separation portion into low-pressure refrigerant; an exterior heat exchanger in which the low-pressure refrigerant flowing out of the lower-pressure side expansion device evaporates, the exterior heat exchanger causing the evaporated low-pressure refrigerant to flow to the suction port; and an auxiliary heater which heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.
 3. A refrigerant cycle device comprising: a compressor having a suction port through which low-pressure refrigerant is drawn, a discharge port through which high-pressure refrigerant compressed in a compression portion is discharged, and an intermediate pressure port through which intermediate-pressure gas refrigerant is drawn to be combined with refrigerant being compressed in the compression portion; a using-side heat exchanger which heats a heat-exchange fluid by performing heat exchange between the heat-exchange fluid and the high-pressure refrigerant discharged from the compressor; a refrigerant branch portion at which a refrigerant passage of the high-pressure refrigerant flowing out of the using-side heat exchanger branches into a first refrigerant passage and a second refrigerant passage, a first expansion device provided in the first refrigerant passage to decompress the high-pressure refrigerant flowing out of the using-side heat exchanger into intermediate-pressure refrigerant; an inner heat exchanger in which the high-pressure refrigerant flowing from the using-side heat exchanger through the second refrigerant passage exchanges heat with the intermediate-pressure refrigerant decompressed by the first expansion device, the inner heat exchanger causing the heat-exchanged intermediate-pressure refrigerant to flow to the intermediate pressure port; a second expansion device configured to decompress the heat-exchanged high-pressure refrigerant flowing out of the inner heat exchanger into low-pressure refrigerant; an exterior heat exchanger in which the low-pressure refrigerant flowing out of the second expansion device evaporates, the exterior heat exchanger causing the evaporated low-pressure refrigerant to flow to the suction port; an auxiliary heater which heats the heat-exchange fluid before or at the same time as that the using-side heat exchanger heats the heat-exchange fluid.
 4. The refrigerant cycle device according to claim 1, wherein the auxiliary heater is arranged upstream of the using-side heat exchanger in a flow direction of the heat-exchange fluid to heat the heat-exchange fluid before the heat-exchange fluid being heated by the using-side heat exchanger.
 5. The refrigerant cycle device according to claim 1, wherein the auxiliary heater and the using-side heat exchanger are arranged in a direction perpendicular to a flow direction of the heat-exchange fluid, and are integrated with each other to heat the heat-exchange fluid at the same time.
 6. The refrigerant cycle device according to claim 1, wherein the auxiliary heater has a heating capacity lower than a standard heating capacity, which is defined as a necessary heating capacity of the auxiliary heater for heating the heat-exchange fluid to a target temperature in a case where (i) the auxiliary heater is arranged to heat the heat-exchange fluid which has been heated in the using-side heat exchanger, and (ii) the heating capacity of the auxiliary heater and a heating capacity of the using-side heat exchanger are used for heating the heat-exchange fluid.
 7. The refrigerant cycle device according to claim 1, further comprising a heating capacity adjusting portion configured to adjust a capacity of the auxiliary heater for heating the heat-exchange fluid so that a pressure of refrigerant in the using-side heat exchanger becomes a target pressure.
 8. The refrigerant cycle device according to claim 1, wherein the auxiliary heater is an electric heater which generates heat by receiving supply of electric power.
 9. The refrigerant cycle device according to claim 1, wherein the auxiliary heater is an auxiliary heat exchanger configured to heat the heat-exchange fluid by utilizing heat of heat medium that cools an external heat source.
 10. The refrigerant cycle device according to claim 7, wherein the auxiliary heater is an electric heater which generates heat by receiving supply of electric power, and the heating capacity adjusting portion adjusts the heating capacity of the electric heater by adjusting a supply of the electric power to the electric heater.
 11. The refrigerant cycle device according to claim 7, wherein the auxiliary heater is an auxiliary heat exchanger configured to heat the heat-exchange fluid by using heat medium, which cools an external heat source, as a heat source, and the heating capacity adjusting portion adjusts the heating capacity of the auxiliary heat exchanger by adjusting a flow amount of the heat medium flowing into the auxiliary heat exchanger.
 12. The refrigerant cycle device according to claim 7, wherein the auxiliary heater is a plurality of electric heaters configured to generate heat by energization thereof, and the heating capacity adjusting portion adjusts the heating capacity of the plurality of electric heaters by changing number of electric heaters energized.
 13. The refrigerant cycle device according to claim 7 wherein the heating capacity adjusting portion activates the auxiliary heater when the capacity of the using-side heat exchanger for heating the heat-exchange fluid is incapable of being increased by a heating capacity control of the using-side heat exchanger.
 14. The refrigerant cycle device according to claim 7, wherein the heating capacity adjusting portion decreases the capacity of the auxiliary heater for heating the heat-exchange fluid when the pressure of refrigerant in the using-side heat exchanger is higher than the target pressure, and the heating capacity adjusting portion increases the capacity of the auxiliary heater for heating the heat-exchange fluid when the pressure of refrigerant in the using-side heat exchanger is equal to or lower than the target pressure.
 15. The refrigerant cycle device according to claim 2, wherein the auxiliary heater is arranged upstream of the using-side heat exchanger in a flow direction of the heat-exchange fluid to heat the heat-exchange fluid before the using-side heat exchanger heats the heat-exchange fluid.
 16. The refrigerant cycle device according to claim 2, wherein the auxiliary heater and the using-side heat exchanger are arranged in a direction perpendicular to a flow direction of the heat-exchange fluid, and are integrated with each other to heat the heat-exchange fluid at the same time.
 17. The refrigerant cycle device according to claim 2, wherein the auxiliary heater has a heating capacity lower than a standard heating capacity, which is defined as a necessary heating capacity of the auxiliary heater for heating the heat-exchange fluid to a target temperature in a case where (i) the auxiliary heater is arranged to heat the heat-exchange fluid which has been heated in the using-side heat exchanger, and (ii) the heating capacity of the auxiliary heater and a heating capacity of the using-side heat exchanger are used for heating the heat-exchange fluid.
 18. The refrigerant cycle device according to claim 2, further comprising a heating capacity adjusting portion configured to adjust a capacity of the auxiliary heater for heating the heat-exchange fluid so that a pressure of refrigerant in the using-side heat exchanger becomes a target pressure.
 19. The refrigerant cycle device according to claim 2, wherein the auxiliary heater is an electric heater which generates heat by receiving supply of electric power.
 20. The refrigerant cycle device according to claim 2, wherein the auxiliary heater is an auxiliary heat exchanger configured to heat the heat-exchange fluid by utilizing heat of heat medium that cools an external heat source. 